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STEAM POWER 






BY 

C. F. HIRSHFELD, M.M.E. 

PROFESSOR OF POWER ENGINEERING, SIBLEY COLLEGE, CORNELL UNIVERSITY 

AND 

T. C. ULBRICHT, M.E., M.M.E. 

FORMERLY INSTRUCTOR, DEPARTMENT (>F POWER ENGINEERING, SIBLEY COLLEGB 

CORNELL UNIVERSITY; ASSOCIATE MEMBER AMERICAN 

SOCIETY OF MECHANICAL ENGINEERS 



FIRST EDITION 

TOTAL ISSUE, SIX THOUSAND 



NEW YORK 

JOHN WILEY & SONS, Inc. 

London: CHAPMAN & HALL, Limited 

1916 




I 



COPYKIGHT, 1910, BY 

C. P. HIRSHFELD and T. C. ULBRICHT 



Ligrary of Congreaa 
By transfer from 
War Department. 

NIAH 29 



BRAUNWORTH & CO. 

BOOK MANUFACTURERS 

BROOKLYN. N. Y. 



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PREFACE 



The following pages represent the results of an attempt 
to collect in a comparatively small book such parts of the 
field of steam power as should be familiar to engineers 
whose work does not require that they be conversant with 
the more complicated thermodynamic principles considered 
in advanced treatments. The experience of the authors 
has led them to believe that a book of this sort should give 
a correct view-point with regard to the use of heat in the 
power plant even though it does not enter deeply into the 
theoretical considerations leading up to that view-point; 
that it should supply the tools required for the solution of 
power plant problems of the common sort; and that it should 
give sufficient description of power plant apparatus to 
make the reader fairly familiar with the more common 
types. 

Mathematical treatment of the subject has been elim- 
inated to the greatest possible extent, and anyone familiar 
with elementary algebra should be able to understand 
readily such equations as it has been deemed necessary to 
include. 

Brief explanations of physical and chemical concepts 
are given in every case in which the text required their use, 
so that those who have not studied these subjects, and those 
who have but have failed to crystallize and hold the neces- 



iv PREFACE 

sary ideas, should have little difficulty in reading the text 
understandingly. 

It is hoped that the book may prove serviceable as a 
text for steam power courses given to civil engineers in the 
various colleges and that it may also meet the needs of those 
instructing power plant operators in industrial schools. 

C. F. H. 

T. C. U. 

June, 1916. 



CONTENTS 



CHAPTER I 

PAGE 

Physical Conceptions and Units 1 

1. Matter. 2. Energy. 3. Units of Matter and Energy. 
4. Work. 5. Mechanical Energy." 6. Heat. 7. Temper- 
ature. 8. Measurement of Temperature. 9. The Unit of 
Heat Energy. 10. Specific Heat. 11. Quantity of Heat. 
12. Work and Power. 

CHAPTER II 

The Heat-power Plant 20 

13. The Simple Steam-power Plant. 14. Cycle of Events. 
15. Action of Steam in the Cylinder. 16. Hydraulic Analogy. 

CHAPTER III 

Steam 27 

17. Vapors and Gases. 18. Properties of Steam. 19. 
Generation of Steam or Water Vapor. 20. Heat of Liquid, 
q or h. 21. Latent Heat of Vaporization, r or L. 22. Total 
Heat of Dry Saturated Steam, A or H. 23. Total Heat of Wet 
Steam. 24. Heat of Superheat. 25. Total Heat of Super- 
heated Steam. 26. Specific Volume of Dry Saturated 
Steam, V or S. 27. Specific Density of Dry Saturated 

Steam, — or 8. 28. Reversal of the Phenomena Just De- 
scribed. 29. Generation of Steam in Real Steam Boiler. 
30. Gauge Pressure. 

CHAPTER IV 

The Ideal Steam Engine 43 

31. The Engine. 32. Operation of the Engine. 33. Work 
Done by the Engine. 34. Heat Quantities Involved. 35. 
Efficiency. 36. Effect of Wet Steam. 37. Application to 



vi CONTENTS 

PAGE 

a Real Engine. 38. Desirability of Other Cycles. 39. The 
Complete Expansion Cycle. 40. The Incomplete Expan- 
sion Cycle. 

CHAPTER V 

Entropy Diagram 61 

41. Definitions. 42. Temperature-Entropy Chart for 
Steam. 43. Quality from TV-Chart. 44. Volume from T^- 
chart. 45. Heat from T0-chart. 46. The Complete T<£- 
chart for Steam. 

CHAPTER VI 

Temperature Entropy Diagrams of Steam Cycles 72 

47. Complete-expansion Cycle. 48. Area of Cycle Repre- 
sentative of Work. 49. Modifications for Wet and Super- 
heated Steam. 50. Incomplete Expansion Cycle. 51. 
Effect of Temperature Range on Efficiency. 

CHAPTER VTI 

The Real Steam Engine 77 

52. Operations of Real Engine. 53. Losses in Real In- . 
stallations. 54. Clearance. 55. Cushion Steam and Cylinder 
Feed. 56. Determination of Initial Condensation. 57. 
Methods of Decreasing Cylinder Condensation. 58. Classi- 
fication of Steam Engines. 59. Rotative Speeds and Piston 
Speeds. 60. The Simple D-Slide Valve Engine. 61. Engine 
Nomenclature. 62. Principal Parts of Engines. 

CHAPTER VIII 

The Indicator Diagram and Derived Values 115 

63. The Indicator. 64. Determination of I.h.p. 65. 
Conventional Diagram and Card Factors. 66. Ratio of 
Expansion. 67. Determination of Clearance Volume from 
Diagram. 68. Diagram Water Rate. 69. T^-diagram for 
a Real Engine. 70. Mechanical and Thermal Efficiencies. 

CHAPTER IX 

Compounding 141 

71. Gain by Expansion. 72. Compounding. 73. The 
Compound Engine. 74. Cylinder Ratios. 75. Indicator 
Diagrams and Mean Pressures. 76. Combined Indicator 
Diagrams. 



CONTENTS vii 

CHAPTER X 

PAGE 

The D-Slide Valve 159 

77. Description and Method of Operation. 78. Steam Lap. 
79. Lead. 80. Angle of Advance. 81. Exhaust Lap. 82. 
The Bilgram Diagram. 83. Exhaust and Compression. 84. 
Diagram for Both Cylinder Ends. 85. Piston Positions. 
86. Indicator Diagram from Bilgram Diagram. 87. Limita- 
tions of the D-slide Valve. 88. Reversing Engines. 89. Valve 
Setting. 

CHAPTER XI 

Corliss and Other High-efficiency Engines 196 

90. The Trip-cut-off Corliss Engine. 91. Non-detaching 
Corliss Gears. 92. Poppet Valves. 93. The Una-flow En- 
gine. 94. The Locomobile Type. 

CHAPTER XII 

Regulation 213 

95. Kinds of Regulation. 96. Governor Regulation. 97. 
Methods of Varying Mean Effective Pressure. 98. Con- 
stant Speed Governing. 99. Governors. 

CHAPTER XIII 

The Steam Turbine 221 

100. The Impulse Turbine. 101. Theoretical Cycle of 
Steam Turbine. 102. Nozzle Design. 103. Action of Steam 
on Impulse Blades. 104. De Laval Impulse Turbine. 105. 
Gearing and Staging. 106. The Reaction Type. 107. Com- 
bined Types. 108. Economy of Steam Turbines. 

CHAPTER XIV 

Condensers and Related Apparatus 251 

109. The Advantage of Condensers. 110. Measurement 
of Vacuums. 111. Conversion of Readings from Inches of 
Mercury to Pounds per Square Inch. 112. Principle of the 
Condenser. 113. Types of Condensers. 114. The Jet Con- 
denser. 115. Non-Contact Condensers. 116. Water Re- 
quired by Contact Condensers. 117. Weight of Water 
Required by Non-contact Condensers. 118. Relative Ad- 
vantages of Contact and Surface Condensers. 119. Cool- 
ing Towers. 



viii CONTENTS 

CHAPTER XV 

PAGE 

Combustion 277 

120. Definitions. 121. Combustion of Carbon. 122. 
Combustion to CO. 123. Combustion to C0 2 . 124. Com- 
bustion of CO to C0 2 . 125. Conditions Determining Forma- 
tion of CO and C0 2 . 126. Flue Gases from Combustion of 
Carbon. 127. Combustion of Hydrogen. 128. Combustion 
of Hydrocarbons. 129. Combustion of Sulphur. 130. Com- 
bustion of Mixtures. 131. Temperature of Combustion. 

CHAPTER XVI 

Fuels 296 

132. Commercial Fuels. 133. Coal. 134. Coal Analyses. 
135. Calorific Value of Coals. 136. Purchase of Coal on 
Analysis. 137. Petroleum. 

CHAPTER XVII 

Steam Boilers 305 

138. Definitions and Classifications. 139. Functions of 
Parts. 140. Furnaces and Combustion. 141. Hand Firing. 
142. Mechanical Grates. 143. Smoke and its Prevention. 
144. Mechanical Stokers. 145. Rate of Combustion. 146. 
Strength and Safety of Boiler. 147. Circulation in Boilers. 
148. Types of Boilers. 149. Boiler Rating. 150. Boiler Effi- 
ciencies. 151. Effects of Soot and Scale. 152. Scale. 153. 
Scale Prevention. 154. Superheaters. 155. Draft Apparatus. 

CHAPTER XVIII 

Recovery of Waste Heat 375 

156. Waste Heat in Steam Plant. 157. Utilization of 
Exhaust for Heating Buildings. 158. Feed-water Heating. 

CHAPTER XIX 

BOILER-FEED PUMPS AND OTHER AUXILIARIES 382 

159. Boiler-feed Pumps. 160. The Steam Injector. 161. 
Separators. 162. Steam Traps. 163. Steam Piping. 



STEAM POWER 



CHAPTER I 
PHYSICAL CONCEPTIONS AND UNITS 

1. Matter. The universe is generally pictured as com- 
posed of matter and energy. Matter is regarded as that 
which is possessed of mass, or as that which is possessed of 
inertia; i.e., which requires the action of force to put it in 
motion, to bring it to rest or to change its velocity. These 
definitions merely enumerate characteristics of matter; they 
do not tell what it really is. In the present state of knowledge 
it is, however, impossible to define matter in any other way. 

No experiment has yet shown that matter can be created 
or destroyed by man. It can be changed from one form to 
another, it can be given certain physical and certain chem- 
ical characteristics, more or less at will, but the actual 
quantity of matter concerned is always the same after and 
before such changes. It is customary to state this experi- 
ence in the form of a law known as the Law of the Con- 
servation of Matter, which states that the " total quantity 
of matter in the universe is constant" 

Matter is known to exist in several physical states or 
conditions of aggregation. The three most familiar are (1) 
solid, (2) liquid and (3) gaseous. In each of these states 
matter is conceived as made up of minute particles called 
molecules which in turn are apparently composed of still 
smaller parts known as atoms. These atoms can also be 
broken into parts, but for the purposes of this book it is not 
necessary to consider such divisions. 



2 STEAM POWER 

Experiment and mathematical reasoning seem to indi- 
cate that the molecules of all materials are in constant 
motion and that there are neutralizing attractive and repul- 
sive forces acting between them. In solids the molecules are 
apparently bound together in such a way that, although they 
are in constant motion, the external form or shape of the 
body tends to remain constant; in fact it requires the 
expenditure of force to cause a change of form. In liquids 
the molecular attraction is so altered that practically all 
rigidity disappears and the shape assumed by the liquid is 
determined by that of the surrounding surfaces, as, for 
instance, the shape of the vessel containing the liquid. In 
gases the molecules are still more free and actually tend to 
move apart as far as possible, so that a gas will spread in 
all directions until it fills any closed containing vessel. 

2. Energy. Nearly everyone has a conception of what 
is meant by the term energy, but no one yet knows what 
energy really is. It is defined as the capacity for doing work, 
or the ability to overcome resistance. A man is said to be very 
energetic or to be possessed of a great deal of energy when 
he has the ability to perform a great amount of work or 
to overcome' great resistances. Matter is said to be pos- 
sessed of energy when it can perform work or overcome 
resistance. Actually, matter is not known in any form in 
which it is not possessed of energy. 

There are many different forms of energy. A body in 
motion can do work and is said to be possessed of mechani- 
cal energy. A body which we recognize as hot can do work 
at the expense of the heat associated with it and is said 
to be possessed of heat energy. Light, sound and electricity 
are all forms of energy. 

Experiment and experience have never shown that energy 
can be destroyed or created by man, but they have shown 
that one form of energy can be converted into another form 
under proper conditions. The first part of this experience 
is stated as a law known as the Law of the Conservation of 



PHYSICAL CONCEPTIONS AND UNITS 



Energy. This law states that " the total quantity of energy 
in the universe is constant" 

3. Units of Matter and of Energy. When attempts are 
made to measure the amount of anything, some unit of 
measurement is adopted. Matter is measured in numerous 
ways and many units are used. The common methods of 
measuring matter are by volume and by weight. Engineers 
in English-speaking countries use the cubic yard, the cubic 
foot or the cubic inch as units in measuring matter by vol- 
ume and they use the pound, the ounce, the grain, etc. as 
units in measuring matter by weight. 

Energy is measured in many units and, in general, there 
is a characteristic unit or set of units for each form in which 
it occurs. Thus the foot-pound is very commonly used for 
measuring mechanical energy; the British thermal unit for 
measuring heat energy; and the joule for measuring electrical 
energy. Some of these units will be defined and considered 
in greater detail in subsequent paragraphs. 

4. Work. Work is defined as the overcoming of 
a resistance through a distance. Thus, work is done 
when a weight is raised against the resistance 
offered by gravity; work is done when a spring is 
compressed against the resistance which the 
metal offers to change of shape; work is done 
when a body is moved over another against the 
resistance offered by friction. 

The unit of work is the quantity of work which 
must be done in raising a weight of one pound 
through a vertical distance of one foot. It is called 
the foot-pound. Thus, one foot-pound of work 
must be done in raising one pound one foot; two 
foot-pounds of work must be done in raising two 
pounds one foot or in raising one pound two feet. 

If a weight of one pound were suspended from 
a spring balance as shown in Fig. 1, the balance would in- 
dicate a pull or force of one pound. No work would be 




Fig. 1. 



4 STEAM POWER 

done by this force as long as the weight remained stationary, 
because no resistance would be overcome through a distance. 
If, however, the same weight were slowly or rapidly raised 
a vertical distance of a foot, one foot-pound of work would 
be done. A force or pull of one pound would then have 
overcome a resistance of one pound through a distance of 
one foot. In general: 

Work in ft. -lbs. = Resistance overcome in lbs. X distance. 
= Force in lbs. X distance in ft. 

so that if a force of 10 lbs. pushes or pulls anything which 
offers a resistance of 10 lbs. while that something travels 
a distance of, say, 5 ft., the work done will be given by the 
expression, 

Work = 10X5, 

= 50ft,-lbs. 

A body in falling a certain distance can do work equal to 
its weight multiplied by the distance it falls because it could 
theoretically raise an equal weight an equal distance against 
the action of gravity, and the work done upon this second 
body would be equal to its weight multiplied by the distance 
through which it was raised. 

It is very important to note that no work is done by a 
force if there is no motion; resistance must be overcome 
through a distance in order that work may be done. Thus, 
a force of 1000 lbs. might be required to hold something in 
position, that is to balance a resistance, but no work would 
be done if the body upon which the 1000-pound force acted 
did not move. Again, a weight of 50 lbs. held at a distance of 
10 ft. above the surface of the earth would exert a downward 
push or pull equal to 50 lbs. on whatever held it in that 
position; it would, however, do no work if held in that 
position. If allowed to fall through the distance of 10 ft. 
it could do 50X10 = 500 ft.-lbs. of work. 

It is very convenient to represent graphically the action 



PHYSICAL CONCEPTIONS AND UNITS 5 

of forces overcoming resistances, that is, doing work. This 
is done by plotting points showing the magnitude of the 
force at the time that the body on which it is acting has 
traveled different distances. Thus, suppose a constant 
force of 10 lbs. pushes a body a distance of 15 ft. against a 
constant resistance of 10 lbs. The force acting on the body 
will have a value of 10 lbs. just as the body starts to move, 
a value of 10 lbs. when the body has moved 1 ft., a value 
of 10 lbs. when the body has moved 2 ft., and so on. This 
might be represented by points on squared paper as shown 



12 
11 






























































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9 

■8 8 

a 

O i 

s 6 

o 5 

o 4 
fa 
3 

2 
1 















































































































































































































































































































5 6 7 8 9 10 11 12 13 14 15 
Distance traveled in Feet 

Fig. 2. 



in Fig. 2 or by a horizontal line joining those points as shown 
in the same figure. 

The work done by this force would be 10X15 = 150 
ft.-lbs. according to our previous definition. But 10X15 
is also the number of small squares under the line represent- 
ing the action of this force in Fig. 2. The number of these 
small squares then must be a measure of the work done, 
but it is also a measure of the area under the line represent- 
ing the action of the force, so that this area must be a measure 
of the work done. Each small square represents 1 lb. by 
its vertical dimension and l.ft. by its horizontal dimension, 



6 STEAM POWER 

so that its area must represent 1 lb.Xl ft. = 1 ft. -lb. 
The total number of squares below the line equals 10 X 
15 = 150, and since the area of each one represents 1 ft.-lb. 
the total area under the line represents 150X1 = 150 
ft.-lbs. 

It is not always convenient to choose such simple scales 
as those just used. Thus it might be more convenient 
to plot the action of this force as is done in Fig. 3. Here 
the height of a square represents 2 lbs. and the width 
represents 1 ft.; the area then represents 2X1=2 ft.-lbs. 
There are 5X15 = 75 squares under the line and as each 



,„ 12 

to 

lio 

1* 

.5 6 






5 6 7 8 9 10 11 
Distance traveled in Feet 



12 13 . 14 15 



Fig. 3. 



represents 2 ft.-lbs. the total area under the line represents 
2X75 = 150 ft.-lbs. as before. 

This is a very useful property of these diagrams and the 
area under the line representing the action of the force 
always represents the work done, no matter what the shape 
of that line. 

Thus, assume a force which compresses a spring a distance 
of 6 ins. Suppose that a force of 10 lbs. is required to com- 
press the spring 1 in., a force of 20 lbs. to compress it 2 ins., 
and so on up to a force of 60 lbs. to compress it 6 ins. 
Starting with a force of zero, the force will have to gradually 
increase as the spring is compressed, as shown by the line in 
Fig. 4. The area of each of the small squares will represent 

10Xt^ = t^ ft.-lbs. Under the line there is an area equal 

1 L 1 — 



PHYSICAL CONCEPTIONS AND UNITS 7 

„o ^~= 18 small squares, and the work done in compressing 

A 

the spring must then be 18X^ = 15 ft.-lbs. 

5. Mechanical Energy. Any body which exists in such 
a position or location that it could do work by dropping or 
falling is said to be possessed of potential mechanical energy, 
or of mechanical energy due to position. As long as it 
remains in this position, it cannot do work at the expense 
of this energy, but, if allowed to fall, it could do so. The 



m 

50 

c 

3 
O 
Cu 

O20 
Cm 

10 











































































Fig. 4. 



l" 



2" 



3" 



y 12 Ft. %, Ft. y n Ft. 

-Graph Showing Action of Spring. 



work it could do would be equal to the product of its weight 
by the distance it could fall and the potential energy it 
possesses before starting to fall is measured by this work. 
Thus, a body weighing 40 lbs. located 10 ft. above the surface 
of the earth could do 40X10 = 400 ft.-lbs. of work in falling, 
and, therefore, it is said to be possessed of 400 ft.-lbs. of 
potential energy before it starts to fall. 

If in falling it raises a weight equal to its own (theo- 
retically) through a distance equal to that through which 
it falls (theoretically), it will have used up 400 ft.-lbs. of 
energy in doing 400 ft.-lbs. of work upon the body raised 



8 STEAM POWER 

and will no longer be possessed of that amount of potential 
energy. The body which has been raised will, however, 
have an equal amount of energy stored in it and will in turn 
be able to do 400 ft. -lbs. of work if allowed to fall a distance 
of 10 ft. 

If the body assumed above falls through a distance of 
10 ft. without raising another body or doing an equivalent 
amount of work in some other way, it acquires a high 
velocity. When it arrives at the bottom of the fall of 10 
ft., it certainly does not possess the 400 ft. -lbs. of potential 
energy which it had before dropping nor has it done work 
at the expense of that energy. Moreover, the energy could 
not have been destroyed because it is indestructible. The 
only conclusion is that it must still be possessed of this 
energy in some way. At the end of the fall it has lost its 
advantageous position, but it has acquired a high velocity, 
and experience shows that if brought to rest it can do 
work upon that which brings it to rest equal to what 
it could have done in raising a weight as previously 
described. 

At the end of its fall and before being brought to rest, 
the body is therefore said to be possessed of energy by virtue 
of its velocity, and this form of energy is called kinetic 
mechanical energy. The kinetic energy will be exactly equal 
to the potential energy which disappeared as the body fell. 

Any body which is moving is possessed of kinetic energy 
because it can do work on anything which brings it to 
rest. This energy is expressed by the equation, 

1 W 

Kinetic JEnergy in ft.-lbs.=-xX7y^XV 2 y 

in which 

W = the weight of the moving body in pounds. 
V = the velocity in ft. per second, and 
32.2 = a gravitational constant commonly represented by g« 



PHYSICAL CONCEPTIONS AND UNITS 9 

6. Heat. One of the most familiar forms of energy is 
heat, which manifests itself to man through the sense of 
touch. In reality every body with which man is familiar 
possesses an unknown amount of heat energy and it is 
assumed that this heat energy is in some way associated 
with the motions and relative positions of the molecules and 
their constituents. 

For this reason heat is often described as molecular 
activity and is regarded as energy stored up in a substance by 
virtue of its molecular condition. Heat energy can be made 
to perform work in ways which will be discussed later and 
this is proof that it is a form of energy and not a material 
substance, as was once supposed. 

Heat is observed and recorded by its effects on matter, 
producing changes in the dimensions or volumes of objects; 
changes of internal stress; changes of state, as ice to water 
and water to steam; changes of temperature; and electrical 
and chemical effects. 

Neglecting certain atomic phenomena not yet well under- 
stood, the probable source of all heat energy appearing on 
the earth is the sun. Heat, however, may be obtained 
from mechanical and electrical energy; from chemical 
changes; from changes of physical state; from the internal 
heat of the earth. 

7. Temperature. Man early realized that under certain 
conditions bodies felt " hotter " than under other conditions 
and gradually came to speak of the " degree of hotness " as 
the temperature of the body. It was later realized that what 
was really measured as the " hotness " or intensity of heat 
or temperature of a body was the ability of that body to trans- 
mit heat to others and that it had no connection with quantity 
of heat. « 

Thus, if the temperature of two adjacent bodies happened 
to be the same, one of them could not lose heat by trans- 
mitting it to the other, but if the temperature of one 
happened to be higher than that of another, the body at 



10 STEAM POWER 

higher temperature would always lose heat to the one at 
lower temperature. 

As a means of measuring temperature certain arbitrary- 
scales have been chosen. The centigrade scale of tempera- 
ture, for instance, is based upon the temperatures of melting 
ice and boiling water under atmospheric pressure. The tem- 
perature difference between boiling water at atmospheric 
pressure and melting ice at atmospheric pressure is arbi- 
trarily called one hundred degrees of temperature, and the 
temperature of the melting ice is called zero, making that 
of the boiling water 100 degrees. 

Any body which has such a temperature that it will not 
give heat to, or take heat from, melting ice is said to be at 
a temperature of zero degrees centigrade, represented as 0° C. 
Similarly, any body in such a condition that it will not give 
heat to or take heat from water boiling under atmospheric 
pressure is said to have a temperature of 100° C. A body 
with a temperature exactly half way between these two 
limits would then be said to have a temperature of 50° C. 

8. Measurement of Temperature. The temperatures of 
bodies could be determined by bringing them in contact with 
such things as melting ice and boiling water and determining 
whether or not a transfer of heat occurred, but this would 
be a very cumbersome and unsatisfactory method. As a 
consequence many other means have been devised for the 
measurement of temperature. 

One of the most common and convenient methods in- 
volves the use of what are known as mercury thermometers. 
These depend upon the fact that the expansion of mercury 
with changing temperature is very uniform over a wide 
temperature range. Thus, if mercury expands a certain 
amount when its temperature is raised from that of melting 
ice to that of boiling water, i.e., 100° C, it will expand just 
half as much when its temperature is raised half as high, 
and one-quarter as much when its temperature is raised one- 
quarter of the range from 0° to 100° C. 



PHYSICAL CONCEPTIONS AND UNITS 



11 



8 



The thermometer is made by enclosing a small quan- 
tity of mercury in a glass tube fitted with a bulb at 
one end, as shown in Fig. 5. The lower end 
of the thermometer is immersed in melting 
ice and the point on the stem which is 
reached by the top of the mercury column is 
marked and labelled 0° C. The thermometer 
is then immersed in the steam from water boil- 
ing under atmospheric pressure and the point 
reached by the top of the mercury column is 
marked and labelled 100° C. The distance be- 
tween the two marks is then divided into one 
hundred parts and each represents the distance 
which the end of the column of mercury will 
move when its temperature changes one centi- 
grade degree. 

It is customary to extend this same scale 
below 0° and above 100°, carrying it, on ex- 
pensive thermometers, as far in each direction 
as the approximation to a constant expansion 
on the part of the mercury and to constant 
properties of the glass justifies. 

The temperature of a body can then 
be found by placing the thermometer in 
or in contact with that body and noting 
the point reached by the end of the 
mercury column. The division reached 
gives the temperature directly. 

The centigrade scale just described is 
the one commonly used by scientists the 
world over, but engineers in this country 

more often use what is known as the 
Fig. (^-Comparison Fahrenheit scale . This is g0 chogen thftt 

the temperature of melting ice is called 
32° F. and the temperature of water boil- 
ing under atmospheric pressure is called 212° F. There are 



Fig. 5.— Mer- 
cury Ther- 
mometer. 



100" 



# # 



of Centigrade and 
Fahrenheit Scales. 



12 



STEAM POWER 



thus 180° on this scale for the same temperature difference 
as is represented by 100° on the centigrade scale. The 
relation between the two scales is shown diagrammatically 
in Fig. 6. It is apparent that the temperature of a body 
at 0° C. will be 32° F. and that of a body at 0° F. will be 
-17.8° C. 

Since 100 centigrade degrees are equal to 180 Fahren- 
heit degrees, it follows that 

180 9_ 



1°C 



= -° f 
100 5 



and that 



100 5 

1 F "180 = 9 C ' 



(2) 



Therefore, if t F and t c represent temperatures on the Fahren- 
heit and centigrade scales respectively, 



fc = gfc+32. 



and 



tc = 7;(tF-32) 



(3) 
(4) 



Cent . ,^-Abs. — ^Fahr. 



100' 


373 2 673 


0° 


273° 491.4° 

459.4 P 


o 
-273 


0° o°U 






There is still another temperature scale of great impor- 
tance. It is known as the absolute 
scale and temperatures measured 
on it are spoken of as absolute tern- 
peratures. The zero on this scale 
is located at -273° C. or 273 centi- 
grade degrees below centigrade 
zero, or, what is the same thing, 
at -459.4° F., or 459.4 Fahrenheit 
degrees below Fahrenheit zero. The 
degrees used are either centigrade 
or Fahrenheit, as convenient, so 
that there are absolute tempera- 
tures expressed in centigrade de- 
grees above absolute zero and there are absolute tempera- 



Fig. 7. — -Comparison of Ab- 
solute v and Ordinary 
Temperature Scales. 



PHYSICAL CONCEPTIONS AND UNITS 13 

tures expressed in Fahrenheit degrees above absolute zero. 
The relations between the various scales are shown dia- 
grammatically in Fig. 7. 

It is apparent from this diagram that, 



and that 



TV = fc+460 (approximately) . . . (5) 
TWc+273 (6) 



if 7V and T c represent absolute temperatures and if the 
number 459.4 is rounded out to 460, as is commonly 
done. 

9. The Unit of Heat Energy. The unit used in the 
measurement of heat energy in the United States is the 
British Thermal Unit (abbreviated B.t.u). It is defined as 
the quantity of heat required to raise the temperature of one 
pound of pure water one degree Fahrenheit. In order to 
make the definition very exact it is iiecessar}' to state the 
temperature of the water before the temperature rise occurs, 
because it requires different amounts of heat to raise the 
temperature of a pound of water one degree from differ- 
ent initial temperatures. For ordinary engineering pur- 
poses, however, such refinements generally may be omitted. 

Many experimenters have shown that heat energy and 
mechanical energy are mutually convertible, that is, the one 
can be changed into the other. When such a change occurs 
no energy can be lost since energy is indestructible, and it 
follows that, if one form is changed into the other, there 
must be just as much energy present after the change as 
there was before. 

As the units used in measuring the two forms of energy 
are very* different and as it is often necessary to express 
quantities of energy taking part in such conversions, it is 
desirable to determine the relations between these units. 
This was first accurately done by Joule, who showed that one 
British thermal unit of heat energy resulted from the con- 






14 STEAM POWER 

version of 772 ft. -lbs. of mechanical energy. Later experi- 
menters have shown that the number 778 more nearly 
expresses the truth than does the number 772 and the larger 
value is now known as Joule's Equivalent. 

Expressed mathematically, the relation between the units 
is 

1 B.t.u. = 778ft.-lbs (7) 

1 ft.-lb. = -^B.t.u ( 8 ) 

778 

10. Specific Heat. The specific heat of a substance is 
defined as that quantity of heat which is used up or recovered 
when the temperature of one pound of the material in question 
is raised or lowered one degree. Its numerical value depends 
upon the specific heat of water since the quantity of heat is 
measured in units dependent upon the amount required to 
raise the temperature of water. The specific heat of water 
is, however, very variable, as shown by the values given in 
Table I., and it is therefore evident that exact numerical 
values of specific heats can only be given when the definition 
of the B.t.u. is exactly expressed. 

The specific heats of all real substances vary with tem- 
perature and the values commonly used are either rough 
averages or are those determined by experiments at one 
temperature. For most engineering purposes errors arising 
from this source may, however, be neglected. 

From the definition of specific heat it follows that : 



c ~w(t 2 -t 1 y (9) 



in which 



C = a mean or average specific heat over a range of tem- 
perature from t\ to t2, and 

Q = the heat supplied to raise the temperature of W 
pounds of material from t\ to fo. 



PHYSICAL CONCEPTIONS AND UNITS 



15 



• TABLE I 

Specific Heats of Water.* 

(Value at 55° F. taken as unity) 



Temp. F°. 


Spec. Ht. . 


Temp. F°. 


Spec. Ht. 


20 


1.0168 


350 


1 . 045 


30 


1.0098 


400 


1 


064 


40 


1.0045 


450 


1 


086 


50 


1.0012 


500 


1 


112 


60 


0.9990 


510 


1 


117 


70 


0.9977 


520 


1 


123 


80 


0.9970 


530 


1 


128 


90 


0.9967 


540 


1 


134 


100 


0.9967 


550 


1 


140 


120 


0.9974 


560 


1 


146 


140 


0.9986 


570 


1 


152 


160 


1.0002 


580 


1 


158 


180 


1.0019 


590 


1 


165 


200 


1.0039 


600 


1 


172 


220 


1.007 






240 


1.012 






260 


1.018 






280 


1.023 






300 


1.029 







* Values taken from Marks and Davis, " Steam Tables and Diagrams," p. 68. 

ILLUSTRATIVE PROBLEMS 

1. Given: Sp. ht. of iron=0.113, of aluminum =0.211; Initial 
temp. = 150° F. Temp, range (fe-fc) = 100° F. 

If 1 lb. of iron and 1 lb. of aluminum are cooled through this 
temperature range, how much more heat is lost in one case than 
in the other? 

Qai = TrCai(* 2 -*i)= IX. 211X100 =21.1 B.t.u. 

Q b = WCM -h) =1 X .113X 100 = 11.3 B.t.u. 

Difference 9.8 B.t.u. 

2. If the difference obtained in Prob. 1 were used to heat up 
5 lbs. of silver, with a specific heat equal to 0.057, what would be 
the temperature range through which it would be raised? 

Q =9.8 =5X0.057(^ 2 -h) =0.285(^ 2 -h) 



16 STEAM POWER 

3. If the initial temperature of the silver in Prob. 2 were 150° F, 
what would be the final absolute temperature Fahr.? 

h =fc+34.4° = 150+34.4 = 184° (approximately). 

T 2 =460+184 =644° F. Abs. 

4. 100 lbs. of water in a 20-lb. tank of iron, both at 60° F., 
are placed in salt brine at 0° F. The water becomes ice at 32° F. 
and the temperature of the ice is lowered to 26° F., the brine being 
raised to 26° F. Sp. ht. water = 1.0; Sp. ht. ice =0.5; Sp. ht. 
iron =0.113; Sp. ht. brine =0.8; and 143 B.t.u. per pound of water 
must be removed to convert liquid water at 32° F. to ice at the 
same temperature. What weight of brine is required? 

100[l(60-32) + 143+.5(32 -26)]+20X0.113(60 -26) 
= 17X0.8(26-0) 
IF =840 lbs. of brine. 

11. Quantity of Heat. It is impossible to determine the 
total quantity of heat in or " associated with " sl substance, 
because no means of removing and measuring all the heat 
contained in any real material have ever been devised. 
Since, however, the engineer is concerned with changes of 
heat content rather than with the total amount of heat 
contained, this fact causes him no difficulty. 

For convenience in figuring changes of heat content, 
it is customary to asstmie some arbitrary starting point or 
datum and to call the heat in the material in question zero 
at that point. 

Thus, for example, if it were necessary to figure heat 
changes experienced by a piece of iron weighing 5 lbs. and 
having a specific heat of 0.1138, and the temperature of this 
iron never dropped below 40° F. under the conditions exist- 
ing, this temperature might be taken as an arbitrary starting 
point above which to figure heat contents. If the iron were 
later found at a temperature of 75° F., " the heat content 
above 40° F." would be said to be 

Q = CW (t 2 -h) =0.1138X5(75-40) =2.27 B.t.u.- 



PHYSICAL CONCEPTIONS AND UNITS 17 

This type of formula can only be used when the sub- 
stance does not change its state between the limits of tem- 
perature concerned. In the case of water which might 
change to steam during such a rise of temperature, it might 
be necessary to include other heat quantities in the cal- 
culations, as shown in a later chapter. 

12. Work and Power. Since steam engines are designed 
for the purpose of converting the heat energy contained in 
fuel into mechanical energy which may be used to perform 
work, it will be necessary to consider the units used in 
measuring work and power. 

Work was defined in a previous paragraph as the over- 
coming of a resistci7ice through a distance, by the application 
of a force; that is, a force expressed in pounds, multiplied by 
the distance in feet through which the force acts, gives a product 
expressed in foot-pounds. 

The amount of work performed in a unit of time is termed 

power, which may be defined as the rate of doing work. 

Therefore, 

_. Force X Distance ,. _ N 

Power = —. —-. r. . . . (10) 

lime (mm. or sec.) 

The unit of power used by steam engineers is the horse- 
power, which is equivalent to the performance of 33,000 
ft.-lbs. of work per minute, or 550 ft.-lbs. of work per second, 
or 1,980,000 ft.-lbs. per hour. Therefore, the horse-power 
developed by any mechanism is 

. ft.-lbs. of work per min. /11X 

h - p - = 337»0- — • • • (ID 

Since 33,000 ft.-lbs. of work can be accomplished only by 

the expenditure of 33,000 ft.-lbs. of energy and since one 

B.t.u. of energy is equal to 778 ft.-lbs., it follows that 33,000 

33 000 ' 

ft.-lbs. of work must be the equivalent of ' = 42.41 B.t.u. 

77o 

It is customary to speak of power in terms of horse- 



18 feTEAM POWER 

power-hours. One horse-power-hour means the doing of 
work equivalent to one horse-power for the period of one 
hour, or the doing of work at the rate of 33,000 ft.-lbs. per 
minute for an hour. A horse-power-hour is therefore equiva- 
lent to 33,000X60 = 1,980,000 ft.-lbs. As 33,000 ft.-lbs. 
are equivalent to 42.41 B.t.u., it follows that 42.41X60 = 
2544.6 or about 2545 B.t.u. are the equivalent of one horse- 
power-hour. 

The number 2545 should be memorized as it is very often 
used in steam-power calculations. If an engine could deliver 
one horse-power-hour for every 2545 B.t.u. it received, it 
would be working without losses of any kind; that is, all 
the heat energy entering it would leave it in the form of 
useful mechanical energy. It will be shown later that this 
is impossible even in the most perfect or ideal engine. 

REVIEW PROBLEMS 

1. Express 32° F. in degrees centigrade. 

2. Express 150° F. in degrees centigrade. 

3. Express 250° C. in degrees Fahrenheit. 

4. Express the results of problems 1, 2 and 3 in absolute values. 

5. What is the heat equivalent of 233,400 ft.-lbs. of work? 

6. Find the heat supplied 10 lbs. of water when its temperature 
is raised from 20° F. to 160° F., assuming the mean specific heat 
over this range to be 0.997. 

7. Find the temperature change of 2 lbs. of lead (sp. ht. 0.0314) 
when 20 B.t.u. are added. 

8. How many B.t.u. must be abstracted to lower the tem- 
perature of 15 lbs. of water from 212° F. to 32° F., assuming the 
specific heat of water to be unity? 

9. Find the weight of water which will have its temperature 
tripled in value by the addition of 250 B.t.u., the final temperature 
being 150° F. Assume specific heat unity. 

10. The specific heat of a piece of wrought iron is 0.113 and of 
a given weight of water is 1.015. 1 cu. ft. of water weighs approxi- 
mately 62.5 lbs. Find the increase in temperature of 4 cu. ft. of 
water when a common temperature of 65° F. results from placing 
in the water a piece of iron weighing 15 lbs. at a temperature of 
900° F. 



PHYSICAL CONCEPTIONS AND UNITS 10 

11. Find the final temperature of the mixture, when 100 lbs. 
of iron (sp. ht. =0.113), at a temperature of 1200° F. are immersed 
in 300 lbs. of water (sp. ht. 1.001) at a temperature of 50° F. 

12. Five pounds of silver (sp. ht. =0.057) at 800° F. are im- 
mersed in water at 60° F., resulting in a final temperature of 85° F. 
Assume Sp. ht. water = 1. What weight of water is necessary? 

13. An engine is developing 10 horse-power. Express this in 
ft .-lbs. of work done per minute and find the amount of heat 
energy equivalent to this quantity of mechanical energy. 

14. A pump raises 1000 lbs. of water 50 ft. every minute. 
How much work is done? Find the equivalent horse-power. 

15. An engine develops 1,980,000 ft.-lbs. of work at the fly- 
wheel per minute. 

(a) Find the horse-power developed. 

(6) If this engine operated in this way for an hour, how many 
horse-power hours would it make available? 

(c) What would be the equivalent of this number of horse- 
power hours in British thermal units? 



CHAPTER II 
THE HEAT-POWER PLANT 

13. The Simple Steam-Power Plant. The various pieces 
of apparatus necessary for the proper conversion of heat 
energy into mechanical power constitute what may be 
called a " Heat-Power Plant," just as the apparatus used 
in obtaining mechanical energy from moving water is called 
an hydraulic or water-power plant. Heat-power plants are 
distinguished as " Steam-Power Plants"; " Gas-Power 
Plants"; etc., according to the way in which the heat of 
the fuel happens to be utilized. 

The apparatus around which the plant as a whole centers, 
that is, the apparatus in which heat energy is received and 
from which mechanical energy is delivered, is termed the 
engine or prime-mover. This heat engine may use steam 
generated in boilers and may require certain apparatus, such 
as condensers, pumps, etc., for proper operation; or it may 
use gas, generated in gas-producers requiring coolers, 
scrubbers, tar extractors and holders, depending upon the 
class of fuel used and upon certain commercial considera- 
tions. Again, the power-plant may simply contain an in- 
ternal-combustion engine using natural gas, gasoline or oil, 
a type of plant which is now very common. 

But whatever type of plant is used, a general method of 
operation is common to all. Heat energy in fuel is constantly 
fed in at one end of the system and mechanical energy is 
delivered at the other end. The steam-power plant will be 
briefly described in the following paragraphs, showing the 
cycle of events with the attendant losses through the 
system. 

20 



THE HEAT-POWER PLANT 



21 




"If;.. | 

I.OT I' 




22 STEAM POWER 

In Fig. 8 is shown a simple steam-power plant which con- 
verts into mechanical energy part of the heat energy, origi- 
nally stored in coal, by means of a prime-mover called a 
steam-engine. The main pieces of apparatus used in this 
type of plant are the steam-boiler; the steam-engine; the 
condenser; the vacuum pump; and the feed-pump. The 
energy stream shows the various losses occurring through- 
out the plant. These losses cause the " delivered energy " 
stream to be only a small fraction of the total heat sent into 
the system. 

14. Cycle of Events. 1. Fuel is charged on the grate 
under the boiler, where it is burned with the liberation of a 
large amount of energy. Air is drawn or forced through 
the grates in proper proportions to support this combustion. 
The hot gases resulting pass over the tubes, in a definite 
path set by the baffle plates, so that the largest possible 
amount of heating surface may be presented to the products 
of combustion. 

There are certain losses accompanying this operation, 
such as radiation, loss of volatile fuel passing off unburned, 
loss of fuel through the grate, and loss of heat through the 
excess air which must always be supplied to insure com- 
bustion. 

2. That part of the heat in the gases which is not lost 
by radiation from the boiler and in the hot gases flowing up 
the stack passes through the heating surfaces of the boiler 
to the water within. From 50 to 80 per cent of the 
total heat energy in the fuel passes through the heating 
surfaces and serves to raise the temperature of the water 
to the boiling point at the pressure maintained, and to con- 
vert this water into steam according to the requirements. 

3. Having obtained steam within the boiler, it is led 
through a system of pipes to a steam engine, where some of 
the heat stored in the steam is converted into mechanical 

• energy by the action of that steam against a piston. The 
steam is then discharged, or exhausted, from the engine 



THE HEAT-POWER PLANT 23 

at a much lower temperature and pressure than when it 
entered. 

From 5 to 22 per cent of the available heat in the 
steam is converted into mechanical energy in the engine 
cylinder, and because of frictional and other losses occurring 
in the mechanism, only from 85 to 95 per cent of this 
energy is turned into useful work at the fly-wheel. 

4. In some plants, known as non-condensing plants, 
the exhaust steam, which still contains the greater part of 
all the heat received in the boiler, is discharged to the atmos- 
phere and represents a complete loss. In others, known as 
condensing plants, the exhaust steam is led to a condenser, 
where it is condensed by cold water, which absorbs and 
carries away the greater quantity of the heat not utilized 
in the engine. The condensed steam or " condensate " is 
then either discharged to the sewer or transferred by means 
of a vacuum-pump to the hot-well, from which it is drawn 
by means of the feed-water-pump, raised to the original 
pressure of the steam, and returned to the boiler. Here 
it is again turned into steam and the cycle of operations 
outlined above is repeated. Naturally there is some loss 
due to evaporation and leaks throughout the system, so 
that " make-up " water must constantly be supplied. 

The series of events just described constitutes a complete, 
closed cycle of operations, wherein the water is heated, 
vaporized, condensed and returned to the boiler, having 
served only as a medium for the transfer of heat energy 
from fuel to engine and the conversion of part of that 
energy within the cylinder. The water in such a case is 
known as the working substance. 

It is often more convenient to discard the working sub- 
stance after it leaves the cylinder, as suggested above in 
the case of a non-condensing plant; or, as in the case of 
a gas engine, where a new supply of working substance 
must be supplied for each cycle, because the burned gases 
of the previous cycle cannot be used again. 



24 STEAM POWER 

15. Action of Steam in the Cylinder. In order to pre- 
pare for the more detailed discussion of the action of the 
steam in the engine cylinder, to be taken up in a later 
chapter, a brief outline of the events occurring within the 
prime-mover will be considered at this point. 

Steam enters the cylinder through some kind of an 
admission valve, and acts upon the piston, just as the latter 
has approximately reached one end of its stroke and is 
ready to return. The heat-energy stored up in the steam 
causes it to expand behind the piston, thereby driving the 
latter out and performing work at the fly-wheel. At about 
90 or 95 per cent of the stroke, the exhaust valve opens, 
and the steam begins to exhaust, the pressure within the 
cylinder dropping almost to atmospheric or to that main- 
ta'ned in the condenser by the time the piston has 
reached the end of its stroke. On the next or return stroke 
the remaining steam is forced out through the exhaust 
port, until, at some point before the end of the piston 
travel, the exhaust valve closes, and the low-pressure steam 
trapped in the cylinder is compressed into the clearance 
space so that its pressure rises. Admission then occurs, and 
the cycle is repeated. 

The diagram given in Fig. 9 illustrates the operation of 

steam within the cylinder. 
This diagram is plotted 
between pressures of steam 
within the cylinder as 
ordinates and correspond- 
ing piston positions as 
abscissas. 
piston Positions The method of obtain- 

Closing of Exhaust Valve 

„ _ „, „ . T ,. ing such a diagram, known 

Fig. 9. — Steam Engine Indicator & ° ' 

Diagram. as an indicator-diagram, 

will be fully described in a 

later chapter. Since vertical ordinates represent pressure 

in pounds per square inch, and horizontal abscissas renre- 





b 

1 

a 


Clearance 

*■ c . ("Jut-off (Olosinp of Admission Valve) 


V. 

c 
p 

u. 

t 


*, ^Sw Release (Opening of 
W '^^"^C^-v Exhaust Valve) 

vS2i>/-« Exhaust --^\*„ 






J~ 1 ~* _ ^Atmo7iVTerToT i r7ssTir7^ r 

Line of Zero Pressures -~» 



THE HEAT-POWER PLANT 



25 



sent feet moved through by the piston, the product of 
these two must be work. But the product of vertical by 
horizontal distances must also give area. Therefore, by 



Source of Water 
at High Head 



Energy Supplied 




Useful Energy 
Made Available 

Energy Discharged 

of Discharged 
Water at Low Head. 



(a) 




Energy Supplied 



Useful Energy 
Made Available 



Energy Discharged 

Receiver of Discharged 
Heat at Low 
Temperature 



(6) 
Fig. 10. — -Hydraulic Analogy 



finding the area enclosed within the bounding lines of the 
cycle and multiplying this by a proper factor, the foot- 
pounds of work developed within the cylinder can be 
determined. 

16. Hydraulic Analogy. The operation of heat-engines 
is analogous to that of water-wheels. A water-wheel de- 



26 STEAM POWER 

velops mechanical energy by receiving water under a high 
head, absorbing some of its energy, and then rejecting the 
fluid under a low head. Similarly, the heat-engine receives 
heat energy at a high temperature (head), absorbs some of 
it by conversion into mechanical energy, and then rejects 
the rest at a low temperature (head). 

The analogy can be carried still further. The water- 
wheel cannot remove all the energy from the water, nor 
can the heat-engine remove all the heat-energy from the 
working substance. There is a certain loss in the material 
discharged in both cases and this cannot be avoided. 

This analogy is illustrated diagrammatically in Fig. 10 
(a) and (b) in which the widths of the streams represent 
quantity of energy. 



i 



CHAPTER III 
STEAM 

17. Vapors and Gases. When a solid is heated, under 
the proper pressure conditions, it ultimately melts or fuses 
and becomes a liquid. The temperature at which this 
occurs depends upon the particular material in question 
and upon the pressure under which it exists. Ice, which 
is merely solid water, melts at 32° F. under atmospheric 
pressure, while iron melts at about 2000° F. under atmos- 
pheric pressure. 

When a liquid is heated, it ultimately becomes a gas, 
similar to the air and other familiar gases. If this gas is 
heated to a very high temperature and if the pressure under 
which it is held is not too great, it very nearly obeys certain 
laws which are simple and which are called the laws of ideal 
gases. 

When the material is in a state between that of a liquid 
and that in which it very nearly obeys the laws of ideal 
gases, it is generally spoken of as a vapor. This term is 
used in several different ways and with several different 
modifying adjectives which will be explained in greater 
detail in later sections. 

18. Properties of Steam. Of the many vapors used by 
the engineer, steam or water vapor is probably the most 
important, because of the ease with which it can be formed 
and also because of the tremendous field in which it can 
be used. It is generated in a vessel known as a steam boiler, 
which is constructed of metal in such a way that it can 
contain water, and that heat energy, liberated from burning 
fuel, can be passed into the water, converting part or all of 
it into water vapor, that is, into steam. 

27 



28 



STEAM POWER 



The properties of water vapor must be thoroughly under- 
stood before the steam engine and steam boiler can be 
studied profitably. Probably the easiest way of becoming 
familiar with these properties is to study the use made of 
heat in the generation of steam from cold water. 

19. Generation of Steam or Water Vapor. For the pur- 
poses of development, assume a vessel of cylindrical form, 
fitted with a frictionless piston of known weight, as shown 
in Fig. 11, (a) and (6), the whole apparatus being placed 
under a bell-jar in which a perfect vacuum is maintained. 



(a) 



jpZglM 




(W 



To Vacuum 
Pump 

Fig. 11. — Formation of Steam at Constant Pressure. 



Assume one pound of water in the cylinder, with the piston 
resting on the surface of the liquid. There will be some 
definite pressure exerted by the piston upon the surface of 
the liquid, and its value will be determined entirely by the 
weight of the piston. 

It is convenient in engineering practice to refer all 
vaporization phenomena to some datum temperature, and 
since the melting point of ice, 32° F., is a convenient refer- 
ence point, it is used as a standard datum temperature, in 
practically all steam-engineering work. Therefore, assuming 
the water in the jar to be at 32° F., if heat is applied the 
temperature of the liquid will rise approximately 1° F. 



STEAM 



29 



for every B.t.u. of heat added, since the specific heat of 
water is approximately unity. 

Experiment shows that for each pressure under which the 
water may exist some definite temperature will be attained at 
which further rise of temperature will cease and the liquid will 



bZi) 
















































/ 






















































/ 






480 














































1 


/ 




















































/ 








440 














































/ 








■g 400 

a 














































/ 


















































) 


' 








|360 

3 












































/ 




















































/ 










f 320 

0> 










































































































P. 

-g 280 

a 






























































































1 












3 
(2240 

l 








































1 


































































3 200 

CO 








































I 


















































1 
















£l60 






































J 


















































I 


















120 




































/ 


















































/ 




















80 
































/ 


' 
















































/ 


/ 






















40 




























/ 


* 














































x* 


1/ 



















































































40 80 120 100 200 240 280 320 300 400 440 480 520 
Temperature -F.° 

Fig. 12.— Pressure-Temperature Relations, Saturated Water Vapor. 



to change to a vapor, that is, to vaporize. The tem- 
peratures at which vaporization occurs at different pressures 
are called the temperatures of vaporization at those pressures. 
The temperatures of vaporization of water are plotted 
against pressure in Fig. 12. It should be noted that the 
values of vaporization temperature increase very rapidly 
for small pressure changes in the case of low pressures, but 



30 STEAM POWER 

that, for the higher pressures, the variation of temperature 
is very small for enormous variations of pressure. This 
fact is of great importance in steam engineering. 

The temperatures of vaporization are tabulated with 
other properties of water vapor in so-called steam tables and 
are constantly referred to by engineers. An example of 
such a table is given on pp. 392 to 399. 

Returning now to the apparatus under discussion, as 
heat is supplied, the temperature of the water will rise from 
32° F. until it reaches the temperature of vaporization cor- 
responding to the pressure exerted upon the water by the 
piston. When this temperature is reached vaporization will 
begin, and if sufficient heat is supplied, will continue without 
change of temperature until the water is entirely converted 
into vapor. 

Up to the time at which vaporization starts the volume 
of the water will change very little, sO that the piston will 
be raised only a negligibly small amount and practically no 
work will be done upon it by the water. On the other hand, 
when vaporization occurs the volume of the material will 
change by a very large amount and the piston will be 
driven out (raised) against the action of gravity. That is, 
work will be done by the steam in driving the piston out 
during the increase in volume which accompanies vaporiza- 
tion. 

It is found that a very great quantity of heat is used up 
during the process of vaporization despite the fact that no 
temperature change occurs. This is described by saying that 
the heat which is supplied during this period becomes latent 
in the steam formed, and the quantity of heat is therefore 
spoken of as the latent heat of vaporization. It is assumed 
to consist of two parts, that used for separating the liquid 
molecules against their attractive forces and that used for 
doing the work which is done upon the piston as it is 
moved upward. The former is called the internal latent 
heat because it is used for doing internal or intermolecular 



STEAM 31 

work ; the latter is called external latent heat because it is 
used for the doing of external work. 

It is to be noted that the internal latent heat may be 
assumed to be tied up in some way within the molecular 
structure of the material and hence to be in the stea?n. 
The external latent heat, on the other hand, is used up as 
fast as supplied for the purpose of driving the piston out 
against the action of gravity. When the piston has been 
raised to any point, the energy used in raising it is not in 
the steam, but is stored as potential energy in the piston. 
To get it back the piston must be allowed to drop. The 
term " external ,} is therefore well chosen; the external 
latent heat is in no sense in the steam; it is stored in 
external bodies or mechanism. 

After the constant temperature vaporization is complete, 
the further addition of heat will again cause a rise of tem- 
perature and a gradual increase of volume. Such raising 
of the temperature of steam already formed is called super- 
heating and results in carrying the vapor nearer and nearer 
to the condition in which it very nearly obeys the laws of 
ideal gases. Since an increase of volume accompanies super- 
heating, the molecules of the vapor must move farther and 
farther apart as superheating progresses. 

Vapor in the condition in which it is formed from the 
liquid and which has the same temperature as the liquid 
from which it was formed is called saturated vapor. This 
term can be pictured as meaning that the maximum number 
of molecules of vapor are packed into a given space; the 
addition of heat to saturated vapor would cause superheating 
and the separation of the molecules so that fewer could be 
contained in a given space. 

20. Heat of Liquid, q or h. Returning once more to 
the start of the process described in the preceding section, 
heat was added to water initially at 32° F. until the tem- 
perature of vaporization corresponding to the existing pres- 
sure was attained. The heat added during this period is 



32 STEAMj POWER 

called the heat of the liquid, and is usually designated by the 
letters q or h. If the mean specific heat of water at con- 
stant pressure (C p ) for the temperature range under con- 
sideration were constant, and equal to 1, then, since 

q = C p (t v -32) 

in which t v is the temperature of vaporization, it would 
follow that, for this pressure 

q = t v -32 (12) 

Therefore, if water boils under a pressure of 50 lbs. at 
a temperature, read from the steam tables, of 281° F., it 
would follow that 

# = 281-32 = 249 B.t.u. 

But the steam tables (see p. 394) for this pressure (50 
lbs.) give 2 = 250.1 B.t.u., indicating, as was shown in 
Chap. I., that the specific heat of water does not remain 
constant, and for this case the mean value must have been 
approximately 1.004 as indicated by the following calcula- 
tion. 

q = C v ft, -32) or 250.1 = C P X 249 
so that 

C,-f£-1.004+ 

Hence it is always advisable to use the steam table 
values of q, except for very approximate calculations. 

21. Latent Heat of Vaporization, r or L. The heat 
supplied during the period of vaporization has already been 
referred to as the latent heat of vaporization and has beer 
divided into internal and external latent heats. 

The internal latent heat is generally designated by p or 
by 7 and the external latent heat by the group of letters 
APu or by E. The group APu merely represents the prod- 



STEAM 33 

uct of pressure, P, by volume change during vaporization, 
u, and by the fraction >lg which is represented by A. The 
product of the first two terms gives external work in foot- 
pounds during vaporization, and dividing this by 778 (Joule's 
Equivalent) converts it to heat units to correspond with the 
other values. It should be noted that P in this expression 
stands for pressure in pounds per square foot. 

The total latent heat of vaporization is generally desig- 
nated by r or by L, and it follows from what has preceded 
that 

r = P +APu (13) 

The value of r for atmospheric pressure, that is, for a 
temperature of vaporization of 212° F., is very often used 
in engineering and should be memorized. Its value is now 
generally taken as 970.4 B.t.u., though recent work would 
seem to indicate a value of about 972 as nearer the truth. 

22. Total Heat of Dry Saturated Steam, X or H. The 
total heat required to convert a pound of water at 32° F. 
into a pound of saturated vapor at some temperature t v is 
called the total heat of the steam or the heat above 32° and 
is designated by X or by H. It is obviously the sum of the 
quantities which have just been considered, so that 

\ = q+r = q + P +APu (14) 

23. Total Heat of Wet Steam. In practical work the 
engineer seldom deals with pure saturated steam, the satu- 
rated vapor nearly always carrying in suspension more or 
less liquid water at its own temperature. To distinguish 
between saturated steam which carries liquid water and that 
which does not, the former is called wet steam or wet 
saturated steam, and the latter dry saturated steam. 

The condition of dryness or wetness is described by what 
is known as the quality of the steam. Dry saturated steam 
is said to have a quality of 100 per cent while saturated 
steam carrying 10 per cent by weight of liquid is said to 



34 STEAM POWER 

have a quality of 90 per cent. Quality expressed as a 
decimal fraction is designated by the letter x, so that if x 
is said to be equal to 0.8 in referring to a certain sample of 
steam, it means that that steam sample consists of 80 per 
cent by weight of saturated steam and 20 per cent liquid 
at the same temperature. 

Since the water in wet steam has the same temperature as 
the stea?n, it contains all the heat of the liquid which it would 
contain if it had been converted into steam, but it obviously 
contains no latent heat of vaporization. It follows that the 
total heat in a pound of wet steam (one pound of a mixture of 
saturated steam and water) with quality equal to x is 

Heat per pound = q+xr = q-\-xp+xAPu . . (15) 

The letter X should never be used in designating the total 
heat per pound of wet steam, as it has been chosen as the 
symbol of the total heat per pound of dry, saturated steam. 

24. Heat of Superheat. When the temperature of 
saturated steam, is raised by the addition of more heat, that 
is, when it is superheated, a very definite quantity of heat 
is required. The quantity required per pound per degreee 
would, by definition, be the specific heat of the material in 
question. 

If the specific heat of superheated steam were reasonably 
constant, the heat required to raise its temperature at 
constant pressure from saturation temperature to some 
higher value fe would be given by the expression 

Heat required per pound = C p (t2 — t v ) 

but superheated steam, as handled by the engineer, is 
generally comparatively near the saturated condition, and 
under these circumstances the values of the specific heat vary 
rapidly with changes of pressure and temperature. The 
extent of these variations is shown in Fig. 13. It will be 
observed that for low pressures the specific heat is approxi- 



STEAM 



35 



mately constant at a value below 0.5 for any given pressure, 
but that for very high pressures it varies widely over a 
comparatively small temperature range. Thus at 600 lbs. 
per square inch the specific heat changes from unity at 
about 510° F. to 0.6 at about 550° F. 

Practically, it is customary to use the type of equation 
just given and to substitute a mean specific heat over the 
required temperature range for the specific heat which can- 
not be assumed constant without too great an error. The 



1.0 



o.8 



^.7 











T_ 












t~ 












T~ 












\ \ 


\ 








^50 


\ \ 


A 






^0 Lbs. 


ier;sq. in. 


' — — 

















100 200 300 400 500 
Temperatures ~Deg. Fahr. 



600 



Fig. 13. — Progressive Valaes of Specific Heat, C P , Water Vapor. 

equation for heat required to raise the temperature from 
U to £2 is then 



Heat of superheat per pound =C pm fe—Q . 



(16) 



in which C pm stands for the mean specific heat at constant 
pressure over the temperature range from t v to fe. 

Values of mean specific heats of superheated steam are 
given in Fig. 14, the values indicated by the curves giving 
the mean specific heat between saturation temperatures and 
various higher temperatures at different pressures. 



36 



STEAM POWER 



25. Total Heat of Superheated Steam. The total heat 
required to convert one pound of water at 32° F. into 
superheated steam at a temperature of fe° F. under constant 
pressure conditions is obviously 

Total heat per ipo\md = q+r+C pm (t 2 — t v ). . (17) 



.70 



.50 



^ ,o 



\\ 


\\ 










V 


\ \ 




V,, 




i 


\ \ 


\\ 


V 


^ 


^ 
^^•s 
^^•/» 




\ 


\\ 




\ 






^ 






^^— 


^: 




->_. 




50 


"~~"^ 














15 


































50 100 150 200 250 

Temperatures above Saturation°F. 



300 



Fig. 14. — -Variation of Mean Specific Heat, Water Vapor. 



and representing the degrees of superheat (t2~t v ) by D, as 
is customary, this becomes 



Total heat per pound = q +r+C pm D. 



(18) 



26. Specific Volume of Dry Saturated Steam, V or S. The 
volume occupied by one pound of a substance is spoken of as 
the specific volume of that material. In the case of dry 
saturated steam there are as many specific volumes as there 
are pressures under which the steam can exist. These 
values are generally tabulated in steam tables and are 
represented by the letter V or the letter S. 



STEAM 



37 



The values of the specific volumes of steam at different 
pressures are given in Fig. 15. It is important to note the 
very gradual change of specific volume at high pressures 
and the very rapid change and enormous increase at low 
pressures. These facts have considerable influence on steam 
engineering practice. 



WV 


•\ 












































480 


























































































440 


























































































«o400 
a 




I 






















































































3 


























































































^320 

0> 


























































































P. 

.§280 


























































































£240 


























































































3 200 

CO 
CO 


















































\ 








































fV.100 






\ 












































\ 


v 






































180 








\ 














































\ 




































80 










s 


s 














































































10 









































































































































2 4 (5 8 10 12 14 16 18 20 22, 
Specific Volume of Dry Saturated Steam (Cubic Feet) 

Fig. 15. — Pressure-Volume Relations, Saturated Water Vapor. 

A curve giving properties of saturated steam is called a 
saturation curve, so that this name may be, and often is, 
applied to the curve given in Fig. 15. 

The volume occupied at any pressure by half a pound 
of dry saturated steam will obviously be half that occupied 
by one pound of such material at the same pressure, and 



38 STEAM POWER 

the same statement can be made for any other fraction of 
a pound. It follows that if the small volume occupied by- 
liquid water in wet steam be neglected, the volume occupied 
by one pound of steam (mixture) of 50 per cent quality 
can be assumed equal to half that occupied by an equal 
weight of dry saturated steam at the same pressure. A 
similar statement could of course be made for any other 
quality and a corresponding fraction. 

Hence if one pound of " wet steam " at a given pressure 
is found to have such a volume that it would be indicated by 
point b in Fig. 16, the quality of this material must be given 

by the expression x = — if the volume occupied by the 

liquid water in the mixture be 
neglected. 

27. Specific Density of Dry 

Saturated Steam, - or 6. The 

weight per cubic foot of saturated 

steam is spoken of as its specific 

density. The specific density 

is obviously the reciprocal of 

the specific volume and is there- 
Fig. 16. — Determining Quality 

from Volume. f ore _ 

28. Reversal of the Phenomena Just Described. If any 

process which has resulted in the absorption of a quantity 
of heat by a substance be carried through in the reverse 
direction, the same amount of heat will again be given up. 
It follows that a pound of dry saturated steam will give up 
the total latent heat of vaporization when condensed to 
liquid at the same temperature, and that the resultant pound 
of hot water will give up the total heat of the liquid if cooled 
to 32° F. 

29. Generation of Steam in Real Steam Boiler. The 
steam boiler is equivalent to a vessel partly filled with water 




STEAM 



39 



and fitted with means for supplying heat to the water and 
for carrying off the vapor formed. This is shown diagram- 
matically in Fig. 17. At first glance this would not seem 
to be at all similar to the cylinder and piston already con- 
sidered, but it really is the exact equivalent so far as the 
generation of steam is concerned. The flow of steam out 
of the steam-pipe is restricted to the extent necessary to 
maintain a high and constant pressure within the boiler, and, 
when in regular operation, steam is formed within the 




Fig. 17. — Formation of Steam in a Steam Boiler. 



boiler under this pressure just as fast as necessary to replace 
that flowing out. 

By picturing the steam as flowing out in layers or lamina 
these lamina can be imagined as taking the place of the 
piston in the apparatus of Fig. 11, and each pound of steam 
formed will then push a piston before it exactly as was 
assumed in the previous discussion. 

30. Gauge Pressure. The steam pressure in a boiler is 
commonly determined by means of an instrument called a 
pressure gauge. These instruments are almost always con- 
structed about as shown in Fig. 18 (a) and (b). The Bourdon 
spring is a tube of elliptical section bent approximately into 



40 



STEAM POWER 



the arc of a circle. One end of this tube is connected directly 
to the pressure connection of the gauge and the other end 
is closed and connected to a toothed sector as shown. 

When the pressure inside a tube of this character is 
increased, the tube has a tendency to unroll or straighten 
out, and in so doing it moves the toothed sector in such a 
way as to rotate the pointer or gauge hand and make its end 
move over the scale in the direction of increasing pressure. 
With diminishing pressure the tube again rolls up and 
rotates the hand in the opposite direction. 




Pressure 
Connection 




(a) (6) 

Fig. 18. — Bourdon Pressure Gauge. 



Instruments of this kind are so made and adjusted that 
the hand points to zero when the gauge is left open to 
the atmosphere. Under such conditions the pressure inside 
the tube is equal to that of the atmosphere and is not zero. 
The gauge therefore only indicates pressures above atmos- 
pheric on its scale, and the total pressure inside the boiler 
is really that shown by the gauge plus that of the atmos- 
phere. 

Pressures as indicated by the gauge are called gauge 
pressures. Pressures obtained by adding the pressure of 



STEAM 41 

the atmosphere to the reading of the gauge are known as 
absolute pressures. Then 

Absolute Pressure = Gauge Pressure -{-Atmospheric Pressure 
and 

Gauge Pressure = Absolute Pressure — Atmospheric Pressure. 

In accurate work the existing atmospheric pressure 
should be determined by means of the barometer, but for 
ordinary, approximate calculations and for cases in which 
no barometric data are available, it is customary to assume 
the pressure of the atmosphere to be equal to 14.7 lbs. per 
square inch. This is very nearly true, on the average, at 
sea level, but is generally far from true at higher elevations. 

PROBLEMS 

1. Determine by means of the steam tables the temperatures, 
total heats, heats of liquid, internal and external latent heats, and 
the specific volumes of 1 lb. of dry, saturated steam under the fol- 
lowing absolute pressures (lbs. per sq. in.): 15, 50, 95, 180 and 400. 

2. Determine the heats of the liquid, latent heats of vapor- 
ization and total heats for 2 lbs. of dry saturated steam at the 
following temperatures in °F.: 101.83, 212 and 327.8. 

3. Determine the volumes occupied by 2 lbs. of dry saturated 
steam under the conditions of problem 2. 

4. Determine the heats of the liquid, latent heats of vapor- 
ization and total heats for 1 lb. of saturated steam with a quality 
of 90% at the following absolute pressures: 25, 50, 75, 125. 

5. Determine the total heat above 32° F. in 12 lbs. of saturated 
steam with quality of 97% at a pressure of 125 lbs. per square inch 
absolute. 

6. What space will be filled by 20 lbs. of dry saturated steam 
at a pressure of 150 lbs. per square inch absolute? 

7. What space will be filled by 20 lbs. of saturated steam at a 
pressure of 150 lbs. per square inch absolute and with a quality 
of 95% if the volume occupied by the water present be neglected? 

8. How many pounds of dry saturated steam at a pressure of 
75 lbs. per square inch absolute will be required to fill a space'of 
10 cu. ft.? 



42 STEAM POWER 

9. How many pounds of saturated steam with quality 96% 
and at a pressure of 110 lbs. per square inch absolute will be required 
to fill a space of 8 cu. ft.? 

10. How much external work, measured in B.t.u., is done when 
1 lb. of water at the temperature of 212° F. is converted into dry 
saturated vapor at the same temperature? 

11. How much external work, measured in foot-pounds, is 
done when 2 lbs. of water at a temperature of 212° F. are converted 
into 90% quality steam at the same temperature? 

12. How much heat is required for doing internal work during 
the vaporization of 1 lb. of water under such conditions that the 
total latent heat of vaporization is 852.7 B.t.u. and the external 
latent heat is 83.3 B.t.u.? 

13. What is the quality of steam containing 1000 B.t.u. above 
32° F. per pound when under a pressure of 150 lbs. per square 
inch absolute? 

14. Heat is added to 1 lb. of mixed steam and water while the 
pressure is maintained constant at 100 lbs. per square inch absolute. 
The percentage of steam in the mixture is increased thereby from 
50% to 95%. 

(a) How much heat was added? 

(b) How much internal latent heat was added? 

(c) How much external latent heat was added? 

15. How much heat is required to completely vaporize 1000 
lbs. of water at a temperature of 92° F. when pumped into a boiler 
in which steam is generated at a pressure of 150 lbs. per square 
inch gauge? Note that heat above 32° F. in 92° F. water is given 
as q in steam tables for a temperature of 92° F. 

16. Find the amount of heat necessary to produce in a boiler 
200 lbs. of steam having a quality of 97% at a pressure of 100 lbs. 
gauge when the feed water has a temperature of 205° F. 

17. What volume would be occupied by the material leaving 
the boiler in problem 16, neglecting volume occupied by water? 



CHAPTER IV 

THE IDEAL STEAM ENGINE 

31. The Engine. If the cylinder and piston assumed in 
the discussion of the last chapter be imagined as turned 
into a horizontal position and fitted with a frame, piston 



-Fly wheel 




Fig. 19. — Simple Steam Engine. 

rod, crosshead, connecting rod, crank shaft and flywheel 
as in Fig. 19, a device results which might be used as a 
steam engine for the production of power. By adding heat 
to, and taking heat from, the water and steam in the cylin- 
der in the proper way and at the proper time, the water 
and steam, or working substance, can be made to do work 
upon the piston. The piston can transmit this work 
through the mechanism to the rim of the flywheel, and it 
can be taken from the rim by a belt connected to a pulley 
on a machine which is to be driven. 

43 



44 



STEAM POWER 



To make the analysis easier, a simplified type of engine 
will be assumed. It is shown in Fig. 20 and consists of the 
same cylinder, piston and piston rod as just described. 
A wire is fastened to the end of the piston rod and run back 
over a pulley in such a way that a weight fastened to the 
free end of the wire will be raised if the piston moves out. 
The weight is made up of two parts, one large and one small. 
When both are on the wire the pull which they exert causes 
the piston to exert a high pressure upon whatever is con- 



a/ 




Volume 

Fig. 20. — Simplified Steam Engine. 



tained in the cylinder. When only the small weight hangs 
on the wire, the piston exerts a much lower pressure upon 
the material in the cylinder. 

Imagine that the piston and the walls of the cylinder 
are made of some ideal material which will not receive or 
conduct heat. Imagine also that the cylinder is fitted 
with a permanent head which is a perfect conductor of 
heat. These conditions are of course ideal but are assumed 
for the sake of simplicity. 

Assume further that, when one pound of water is con- 



THE IDEAL STEAM ENGINE 45 

tained in the cylinder and the piston is driven into the 
cylinder by the two weights until the space between the 
piston and the cylinder head is just large enough to 
contain the pound of water, the piston exerts a high 
pressure equal to Pi pounds per squcre foot against the 
water. The volume of this water and the pressure upon 
it can be represented by the point a of the PV diagram, 
Fig. 20. 

32. Operation of the Engine. With conditions as de- 
scribed in the preceding paragraphs, imagine a flame or 
other source of heat at high temperature to be brought into 
contact with the conducting cylinder head and to pass heat 
into the cylinder, raise the temperature of the water within 
to the temperature of vaporization and ultimately vaporize 
it. As the water vaporizes it will push the piston out of the 
cylinder just as described in the last chapter and a hori- 
zontal line such as ab in Fig. 20 will represent the increase 
of volume (vaporization) at constant pressure. The point 
b may be assumed to represent the volume of one pound 
of dry saturated vapor at a pressure Pi. Obviously the 
steam, as it is formed, does work in driving out the piston 
against the resistance offered by the weights which must be 
raised. 

If a stop is provided which will prevent the movement 
of the piston beyond the position corresponding to the 
point b, it will be possible to remove the larger weight when 
that point is reached and the high pressure steam will hold 
the piston and rod hard against the stop. If now some 
cooling medium is applied, such as a large piece of ice held 
against the conducting head of the cylinder or water running 
over that head, heat will be abstracted and a partial con- 
densation of the steam within the cylinder will occur. 
As condensation progresses the pressure will drop because 
there will be less and less steam, by weight, in a given 
volume. Such a process, would be indicated by the line 
be which represents a drop of pressure, while the volume 



46 STEAM POWER 

contained within the cylinder walls between head and 
piston remains constant. 

When some point c is reached, the steam pressure will 
have been reduced to a value equal to that exerted by the 
small weight, and the piston will be driven in toward the 
cylinder head while the heat absorbing medium continues 
to remove heat from the steam and to cause further conden- 
sation. The combination of piston motion and heat ab- 
sorption will be so regulated that the pressure remains con- 
stant at P2 during this process, because the weight will 
move the piston inward just as fast as necessary to main- 
tain a constant pressure. If sufficient heat is absorbed, 
the pound of material within the cylinder will ultimately 
all be condensed or liquefied and will just fill the volume Va. 

The heat absorbing body may now be removed and an 
infinitesimal motion of the piston toward the head would 
serve to raise the pressure on the liquid water from P2 to 
Pi so that the volume Va may be taken equal to the volume 
V a and the line da may be assumed to be vertical. It 
would then represent an increase of pressure at constant 
volume. This might be caused by hanging a weight of the 
larger size on the wire when condition d was reached. 

Having brought the material, or working substance, 
back to the conditions originally shown at a, the high 
temperature source of heat can again be brought in contact 
with the end of the cylinder and the entire cycle carried 
through once more. There is obviously no reason why it 
could not be repeated as often as desired. 

33. Work Done by the Engine. If the device just 
described is to serve as a steam engine, it must actually 
make mechanical energy available, that is, it must convert 
into mechanical form some of the heat energy supplied it. 
It is now necessary to see whether it does so. 

Water vaporizing and increasing in volume as from 
V a to V b was shown in the last chapter to do work upon the 
piston confining it. Work has been shown to be equal to 



THE IDEAL STEAM ENGINE 47 

(total force X total distance) and in this case if L repre- 
sents the distance in feet traveled by the piston, the work 
done by the steam upon the piston while the latter moves 
from a' to b' must be 

Work done on piston = total force X distance 

= Pi X area of piston XL . . . ft.-lbs. 

But the product of area of piston in square feet by 
distance traveled in feet is equal to the piston displace- 
ment or volume swept through by the piston, that is 
(Vb— Va) cubic feet. Therefore 

Work done on piston = Pi(V b - V a ) ft.-lbs. . (19a) 

Pl(V b -V a ) 



778 



B.t.u. . (196) 



The first form of this expression Pi(V b —V a ) is very 
obviously represented by the area under the line ab in 
Fig. 20 and this area therefore represents the work done by 
the steam upon the piston during the change of volume at 
constant pressure represented by that line. While the 
steam is supplying this amount of energy to the piston or 
doing this amount of work upon the piston, the latter does 
an equivalent amount of work upon the weights if friction- 
less mechanism be assumed. In such a case the total 
weight hung on the wire multiplied by the distance raised 
would therefore give the same result in foot-pounds as 
that just obtained. 

It should be noted that Eq. (19) is merely an expression 
of the external work done during vaporization, that is, 
an expression of the amount of heat which is used for the 
doing of external work. It is the exact equivalent of the 
external latent heat previously discussed. In fact, the 
group of symbols APu is really a condensation of Eq. (19) 

formed by putting A for — — and u for {V b — V a ). 

i to 



48 STEAM POWER 

The line cd also represents a change of volume at con- 
stant pressure and the same type of formula as applied 
to ab will express the work done during this process. In 
this case, however, the piston is being pushed into the 
cylinder by the small weight against the pressure of the 
steam, and energy is being supplied to push the piston in. 
This energy is equal to the weight of the small weight 
(pounds) multiplied by the distance it falls (feet). The 
piston is therefore doing work upon the steam, and the 
amount is 

Work done on steam = P 2 (V C - V d ) ft.-lbs. . . (20) 

P2(Vc-V d ) 



778 



B.t.u. . . (21) 



The first form of expression also represents the area under 
the line cd and this area therefore represents the work 
done by the piston upon the steam mixture in the cylinder 
during the process represented by cd. 

No work can be done by steam on piston or by piston 
on steam during the processes represented by be or da 
because both the weights and the piston are stationary 
during these changes and it has already been shown that 
work involves motion. 

The total work done upon the piston by the steam 
is therefore represented by the area abef and this amount 
of energy is used in raising the two weights through a 
vertical distance equal to the piston travel. Some of 
this energy, or its equivalent, will have to be returned an 
instant later, however, in order that the piston may do the 
work shown by the area cdfe upon the steam. It is returned 
by the small weight dropping through a distance equal to 
the travel of the piston. The net mechanical energy 
made available by carrying through the series of processes 
is therefore represented by the area (abef) — (cdfe) = (abed) 
or the area enclosed by the four lines representing the 



THE IDEAL STEAM ENGINE 49 

pressure and volume changes experienced by the working 
substance during one cycle of events. It is equal to the 
work done in raising the larger weight a vertical distance 
equal to the travel of the piston. 

This net energy made available is obviously 

Energy made available = Pi ( V b — V a ) — P2 ( V c — Va) 

= (Pi-P2)(T r 6 -T r a )ft,-lbs. (22) 

JPl -P 2 W-v a) Btu (23) 

Since this amount of energy is made available while one 
cycle of events is being carried out and since the cycle 
can be repeated time after time if sufficient heating and 
cooling mediums are available, any quantity of mechanical 
energy can be produced from heat energy by repeating the 
cycle a sufficient number of times. This would correspond 
to picking up a number of the larger weights which were 
slid on to the wire at the lower elevation and slid off at 
the higher. 

This repetition of cycles would correspond, in a real 
engine, to running at such a speed that the required number 
of cycles would be produced in a given time to make avail- 
able the amount of mechanical energy required. 

Or, the power made available per cycle could be increased. 
This is easily seen by an inspection of Eq. (22). Increas- 
ing the value of either of the right-hand terms will obviously 
increase the amount of energy made available. The 
value of (P1 — P2) can be increased by raising the initial 
pressure Pi or by lowering the final pressure P2. The value 
of (Vb— V a ) may be increased by using more than one pound 
of material, thus increasing both the volume Vj, of the satu- 
rated steam formed and increasing the volume V a of the 
liquid water, but getting a greater numerical value for 
(Vb—Va). This would correspond in a real case to using 
a larger cylinder and therefore a larger engine. 



50 STEAM POWER 

ILLUSTRATIVE PROBLEM 

An engine of the type described is to work with a maximum 
pressure of 100 lbs. per square inch absolute and a minimum 
pressure of 15 lbs. per square inch absolute. The cylinder is to 
be of such size that 1 lb. of water is used and the steam is to be 
dry and saturated at the point b of the cycle. 

Find: (a) the amount of mechanical energy made available 
per cycle; (b) the amount of energy made available per minute 
if 150 cycles are produced per minute; and (c) the horse power of 
the engine. 

It will first be necessary to find the piston displacement required 
and the space necessary between piston and cylinder head to 
accommodate the pound of water in liquid form. The steam tables 
give the volume of one pound of dry saturated steam at 100 lbs. 
per square inch as 4.429 cu.ft. and the volume of one pound of 
water may be taken as 0.017 cu.ft. The values of the various 
volumes and pressures will therefore be 

Va = V d =0.017 cu.ft.; 

V b ^Vc^ 4.429 cu.ft.; 

p a =p b = 100X 144 = 14,400 lbs. per sq.ft.; 

P c =P d = 15X144 =2160 lbs. per sq.ft. 

(a) Using Eq. (22) the amount of mechanical energy made 
available per cycle will be 

(P, -P 2 )(V b - Va) = (14,400 -2160) (4.429 -0.017) 
= 12,240X4.412; 
= 54,002.88 ft.lbs. 

{b) If 150 cycles are produced per minute, the total amount 
of mechanical energy made available per minute must be 

150X54,002.88=8,100,300 ft.-lbs. 

(c) The horse power must then be 

. 8,100,300 

h - p - = ^ooo- 245+ - 

34. Heat Quantities Involved. It is a very simple 
matter to determine the quantity of heat which must be 
supplied to produce the process ab, and the quantities of 



THE IDEAL STEAM ENGINE 51 

heat which must be removed to produce the processes 
be and cd. This can be done by making use of the known 
properties of water and steam as given in the steam 
tables. 

The water at d must be at the temperature of vaporiza- 
tion corresponding to pressure P2 since it has just been 
formed by condensation from steam under that pressure. 
It therefore contains the heat of the liquid corresponding 
to that pressure. If it is to be vaporized at pressure Pi, it 
must first be raised to the higher temperature corresponding 
to that pressure. The amount of heat required to do this 
will obviously be the difference between the heat of the 
liquid at the temperature corresponding to Pi and the heat 
of the liquid at the temperature corresponding to P2. 
These can be found in the steam tables. 

The latent heat of vaporization at Pi must then be added 
to cause the increase of volume shown by ab. This can 
also be found in the steam tables for any given case. 

The quantity of heat which must be removed to produce 
the processes represented by be and cd can be found sim- 
ilarly from steam table values, although the exact method 
of procedure is not quite as obvious as in the preceding 
cases. 

Assuming that it is possible to find the heat supplied, 
Qi, and the heat removed, Q2, it is obvious that the energy 
made available in mechanical form, per cycle, must be 
equal to (Q1 — Q2) B.t.u., since this is the amount of heat 
energy which has disappeared and since it cannot have 
been destroyed. This may be put in the form of an equa- 
tion, thus 

Energy made available = Qi — Qz. . . (24) 

If the proper substitutions are made in this formula and it 
is then simplified, it becomes 

Energy made available = (APu) Pl — x c (APu) P2 B.t.u., (25) 



52 . STEAM POWER 

in which 

(APu) Pl = the external latent heat at pressure Pi; 
(APu)p 2 = the external latent heat at pressure P2, and 
x c = quality at point c, which can be found from 
the ratio of dc to dc'\ 

Numerical substitution in this equation for any given 
case will show that it gives exactly the same values as would 
be obtained by the use of Eq. (23). 

It is to be noted particularly that the energy made 
available is actually less than the external latent heat at 
the higher pressure, while the heat supplied must be equal 
to the total latent heat plus some of the heat of the liquid. 
An inspection of the steam tables will show that the exter- 
nal latent heat for ordinary steam pressures forms a very 
small fraction of even the total latent heat, and therefore 
the mechanical energy made available for a given expendi- 
ture of heat energy is very small in the case under dis- 
cussion. 

35. Efficiency. The term efficiency is used in engineer- 
ing as a measure of the return obtained for a given expendi- 
ture. It may be defined in any one of the following ways: 

. _ Useful result 

Expenditure made to obtain that result 

_ Result 

"Effort 

=tSt «> 

In the case of a heat engine, the useful result is the 
mechanical energy obtained by the operation of the engine, 
while the expenditure made is the heat which is supplied. 
For this case efficiency may therefore be defined by the 
expression 



THE IDEAL STEAM ENGINE 53 

„ . „ . Mechanical energy obtained per cycle 
Engine efficiency = - Heat supplied per cycle 



Qi' ' 

Q1-Q2 
Qi ' 



(27) 
(28) 



in which 



E stands for mechanical energy obtained, 
Qi stands for heat supplied, and 
Qi stands for heat rejected. 

In the case of the type of steam engine just considered, 
this efficiency would have a value between 6 and 8 per 
cent for ordinary pressures. That is, the engine would 
produce in mechanical form only 6 to 8 per cent of the 
energy supplied it in the form of high temperature heat. 
Moreover, these figures would hold only for a theoretically 
perfect engine; a real engine built to operate upon this 
cycle would probably give efficiencies of the order of 2 to 
3 per cent. The reasons for this great discrepancy will be 
discussed in a later chapter. 

36. Effect of Wet Steam. In what has preceded, it 
was assumed that the pound of steam was completely vapor- 
ized along the line ab so that dry, saturated steam existed 
in the cylinder at b. It might, however, be assumed that 
vaporization was incomplete at the upper right-hand 
corner of the cycle, so that this point occurred at a point 
to the left of b and with a quality x at that point less than 
unity. 

Under such conditions, the cylinder would not have to 
be so big, since the maximum volume attained by the steam 
would be smaller than in the preceding case. The work 
done per cycle would obviously be smaller in quantity, 
because the area enclosed within the lines of the cycle would 
be smaller. It can also be shown that the efficiency would 
also be lowered by lowering the quality at b. 



54 STEAM POWER 

37. Application to a Real Engine. The engine which 
has been described in the preceding paragraphs could easily 
be converted into the counterpart of a real engine by sub- 
stituting connecting rod, crank shaft and flywheel for wire, 
pulley and weights as described in the first paragraph of 
this chapter. It could then be made to do work in just 
the same way as has been described; some of the energy 
made available during the outstroke would be used for 
overcoming resistance at the shaft, that is, doing useful work, 
and some of it would be stored in the flywheel which would 
speed up slightly. The energy which must be expended on 
the steam during the return stroke would be obtained by 
allowing the flywheel to slow down and thus deliver suf- 
ficient kinetic energy to drive the piston back against the 
low-pressure steam. The cycle and the efficiency would 
thus, theoretically, be exactly the same as those just in- 
vestigated . 

Great difficulty would, however, be met in a real engine 
if the steam had to be formed and condensed within the 
cylinder, and another method which gives the same results 
is therefore used. Steam is generated in a boiler and 
allowed to flow into the cylinder and push out the piston 
just as though it were actually being formed in the cylin- 
der as previously described. When the piston reaches the 
end of its outstroke the inlet valve is closed and the exhaust 
valve is opened, allowing some of the steam to blow out 
into a space in which a lower pressure exists. As the 
piston stands still at the end of its stroke while the pres- 
sure drops, the line be is produced as in the previous descrip- 
tion, but by a different method. The piston then returns 
and drives the remaining steam out of the cylinder at a 
constant pressure theoretically equal to that of the space 
into which the steam is being forced or exhausted. The 
line cd is thus produced and the closure of the exhaust 
valve and opening of the admission valve when d is reached 
will start the cycle over again. 



THE IDEAL STEAM ENGINE 55 

In order to get more work out of a given size of cylinder 
and to obviate the necessity of giving back energy which 
has already been given out, engines are generally made to 
take steam on both sides of the piston. They are then 
known as double acting engines. In this case the steam 
admitted on one side of the piston would supply the energy 
necessary both for overcoming the resistance due to the 
load and for driving out the low-pressure steam on the 
other side of the piston. On the return stroke conditions 
would be just reversed. 

38. Desirability of Other Cycles. The cycle of opera- 
tions described in preceding paragraphs is the most inef- 
ficient of all those actually used, that is, it gives the small- 
est return for a given amount of heat supplied. This 
is because only the external latent heat supplied is con- 
verted into mechanical energy and part of that energy 
must be returned to complete the cycle. All of the internal 
latent heat and all of the heat of the liquid supplied along 
ah pass through the engine without conversion and are 
exhausted. 

Therefore, cycles which differ from that described 
in such a way as to make it possible to convert into mechani- 
cal energy some of the internal latent heat and possibly 
some of the heat of the liquid should be highly desirable 
as they ought to yield a larger return of mechanical energy 
for the same total amount of heat supplied. Two such 
cycles are commonly used; they may be described as the 
Complete-expansion cycle and the Incomplete-expansion cycle. 
The former is used in steam turbines, the latter in most 
reciprocating steam engines. The rectangular cycle which 
has just been described is used in duplex pumps and similar 
apparatus. 

39. The Complete-expansion Cycle. This cycle, which 
is also, known as the Clausius and as the Rankine cycle, 
starts just the same as that already described. This is 
shown in Fig. 21. The pressure on, say, a pound of water 



56 



STEAM POWER 



is raised from P2 to Pi and its temperature is raised from 
that of vaporization at P2 to that of vaporization at Pi. 
After this it is vaporized, giving the increase of volume 



CSiTecl Spring b\ 




Volume 
Fig. 21. — Complete Expansion Cycle or Clausius Cycle. 

shown by ab. The supply of heat is then stopped. The 
cylinder of the engine is made larger than in the preceding 
type so that when the point b' is reached the piston can 
travel still further, and it is allowed to do so, that is, the 



THE IDEAL STEAM ENGINE 57 

high-pressure steam is allowed to push it further out. This 
can be pictured by imagining the steam to act like the 
compressed spring shown in the figure and to push the 
piston in much the same way as does the spring. The line 
bici shows the decreasing pressure exerted on the piston by 
the spring as the latter expands so as to get longer and 
longer. Because of the properties of a spring this is a 
straight line. The line be shows the decreasing pressure 
exerted on the piston by the steam as the latter expands 
so as to occupy greater and greater volumes. Because of 
the properties of steam this line is curved instead of straight. 

Work will be done on the piston by the expanding 
steam during the process be and the amount of this work 
will be indicated by the area under the line be as shown 
in the figure. This work must have been done by the 
expenditure of energy on the part of the steam and since 
no energy was added after the point b was reached the work 
must have beeu done at the expense of heat energy contained 
in the steam at b. It has already been shown that the 
heat above 32° in the steam at b is equal to the sum of the 
heat of the liquid and the internal latent heat, and some 
of this heat must obviously be used for the doing of work 
along be instead of being entirely rejected to the cooling 
medium as in the preceding cycle without " expansion." 

The expansion of the steam continues until the " back 
pressure " P2 is reached. The cooling medium may then 
be imagined to be brought into use and to abstract such 
heat of vaporization as may remain in the steam besides 
absorbing the equivalent of the work done on the steam 
by the returning piston, thus giving the process shown by 
the line cd. 

If the expansion line be of the cycle just described 
could, be carried out within walls constructed of such mate- 
rial that it would not give heat to nor take heat from 
the steam, it is obvious that any heat energy lost by the 
steam during the expansion could be lost only by conver- 



58 STEAM POWER 

sion into mechanical energy. An expansion of this kind 
is called an adiabatic expansion. 

In the figure, the curve of adiabatic expansion is shown 
in its correct position with respect to the saturation curve 
and it is obvious that for an adiabatic expansion, starting 
with dry, saturated steam, the quality decreases as the expan- 
sion progresses. 

Comparison with Cycle without Expansion. The heat 
supplied is the same in both of the cycles just considered 
when they operate between the same two pressures, but 
the mechanical energy obtained in the case of the complete 
expansion cycle is much greater. In Fig. 21, for instance, 
the mechanical energy obtainable with the cycle first 
described is represented by the area abd'd while that obtain- 
able with the complete expansion cycle with the same 
heat supply Q\ is represented by the same area abd'd plus 
the additional area bed' . The efficiency of the complete 
expansion cycle is therefore very much higher than that 
of the cycle without expansion. 

For conditions similar to those giving a theoretical 
efficiency of about 6 per cent without expansion, the com- 
plete expansion cycle will give a theoretical efficiency of 
about 12 per cent and this figure can be doubled by 
expedients which will be considered later. 

The cylinder required for the production of the com- 
plete expansion cycle would be much larger than that re- 
quired for the other cycle if both used the same weight of 
steam per cycle. The proportion would be in the ratio 
of the volume shown at c in Fig. 21 to the volume shown 
at b. But the complete expansion cycle would make avail- 
able much more energy per pound of steam than would 
the other, so that the difference in the size of cylinders 
would not be so great if both were required to make avail- 
able the same amount of mechanical energy per cycle. 

40. The Incomplete-expansion Cycle. The shape of 
this cycle is shown in Fig. 22. It is just like the complete 




THE IDEAL STEAM ENGINE 59 

expansion cycle down to the point c. The cylinder in which 
it is produced has a smaller volume than that used for the 
complete expansion cycle so that the piston arrives at the 
end of its stroke before it 

has opened up volume | a b 

enough to enable the 
steam to expand all the 
way down to the lowest 
pressure (terminal or back 
pressure) . When the point 
c is reached in the real 
engine, the exhaust valve 

is opened and enough „ 00 T , , -^ . r , , 

1 ° Pig. 22. — Incomplete .Expansion Cycle. 

steam then blows out to 

reduce the pressure to the back pressure Pa. The piston 
then returns and drives out the remainder of the steam as 
shown by the line de. 

In the ideal method assumed in the preceding treat 
ment, the heat absorbing medium would be brought into 
use at c, absorbing sufficient heat to reduce the pressure 
from P c to Pa while the piston remained stationary at the 
end of its stroke. The latent heat of vaporization remain- 
ing in the steam at d would then be absorbed as the piston 
was driven back from d to e. 

Comparison with Other Cycles. The incomplete expan- 
sion cycle is intermediate between the two previously dis- 
cussed. This can be appreciated readily by an inspection 
of Fig. 22. In this figure the area abd'e represents the 
mechanical energy obtainable with the cycle without 
expansion; the area cibc'e represents the energy obtainable 
from the same quantity of steam with complete expansion; 
and the area abode represents the energy obtainable from 
*-he same amount of steam with incomplete expansion. 

The later the point at which the exhaust valve is opened, 
point c, the more nearly do efficiency and energy obtain- 
able approach the values for the complete expansion cycle< 



60 STEAM POWER 

The earlier the point at which the exhaust valve is opened, 
the more nearly do efficiency and energy obtainable approach 
the values for no expansion. 

Despite the lower efficiency of the incomplete expan- 
sion cycle as brought out in connection with Fig. 22 it is 
universally used on all reciprocating engines excepting 
those which make do pretense to economy and use no 
expansion. The less efficient cycle is used for the simple 
reason that complete expansion in a reciprocating engine 
does not pay commercially. For complete expansion the 
cylinder must be larger in the ratio of V c to V c > as shown 
in Fig. 22 and the work obtained by completing the expan- 
sion is a very small part of the total. In most cases it 
would not be great enough to overcome the friction of the 
engine, not to mention paying interest on the necessarily 
higher cost of the larger cylinder and accompanying parts. 

It will be shown in a later chapter that the steam tur- 
bine can economically expand the steam completely and 
the complete expansion cycle is therefore used with such 
prime movers. 



CHAPTER V 
ENTROPY DIAGRAM 

41. Definitions. In Chapter III temperature, pressure 
and volume were discussed as criteria determining the con- 
dition of water and steam. Other things may be used in 
determining the condition of such materials. One which is 
particularly useful from an engineering standpoint is known 
as entropy and is designated by the Greek letter fa 

For every condition of water and steam, there is a char- 
acteristic value of entropy just as there is a characteristic 
value of temperature, pressure, volume, heat above 32° F., 
etc. These values of entropy are given in the steam tables 
in just the same way as the value of temperature, pressure, 
volume, heat above 32° F., and such, are given. 

The entropy of the liquid given for any particular pres- 
sure is the change of entropy experienced by one pound 
of the liquid when its temperature is raised from 32° F. 
to the temperature of vaporization corresponding to that 
particular pressure. It might be spoken of as the entropy 
of the liquid above 32° F., just as q is spoken of as the heat 
of the liquid above 32° F. It is represented by <f>i. 

The entropy of vaporization given for any particular 
pressure is the change of entropy experienced by one pound 
of the material while changing from water at the tempera- 
ture of vaporization to dry saturated steam at constant 
pressure. It corresponds to the latent heat of vaporiza- 
tion and is designated by fa. 

The entropy of dry saturated steam at any pressure is 
the sum of fa and fa and therefore is the total change of 
entropy experienced by a pound of material in changing 

61 



62 



STEAM POWER 



from water at 32° F. to dry saturated steam at the particu- 
lar pressure in question. 

The entropy of superheat at any pressure and tempera- 
ture is the change of entropy experienced by a pound of 
dry, saturated steam at that pressure when superheated 
to that particular temperature. It is designated by <j> s . 

The entropy of superheated stsam at any pressure and 
temperature is the total change of entropy experienced by 
one pound of material when changed from water at 32° F. 



T>0 



Si 

.1 / \Jf 

£ [) Vaporization Line \7 


4? A c\ Superheat 
S th '\ ivegion 

3 9 i Wet Steam Region ! o,\A 


£l Region of Incomplete] ?X^ 
J | Evaporation "^ \ r> 
/ 1 W/, 
!„ i'A °A.bs. ! C vl 
Entropy Scale 



(a) 



** 




7//\ of Superheat 



Fig. 23. — Temperature-Entropy Diagrams. 



to superheated steam at the pressure and temperature 
in question. It is equal to </>/+<£> +</>*• 

42. Temperature-Entropy Chart for Steam. Entropy 
is particularly useful to the engineer because it enables him 
to draw charts which lend themselves readily to an easy, 
graphical solution of certain problems which would other- 
wise involve complex calculations. One of these charts 
is known as the Temperature-Entropy Chart. 

In making this chart, absolute temperature is generally 
plotted on the vertical and entropy above some datum tem- 
perature on the horizontal, as shown in Fig. 23 (a) and (6), 
which represents the construction of a temperature entropy 
diagram for water and steam. The entropy values on 



ENTROPY DIAGRAM 63 

this chart are plotted above 32° F. as datum tempera- 
ture. 

The water line or water curve is obtained by picking 
out of the steam tables the values of fa, entropy of the liquid, 
for different pressures and plotting them against the abso- 
lute temperatures corresponding to those pressures. Ob- 
viously, zero of entropy will occur at the absolute tempera- 
ture corresponding to 32° F., i.e., about 492° F. abs. 

The saturation curve or dry steam curve is obtained by 
picking out of the steam tables the values of fa-\-fa for 
different pressures and plotting against corresponding 
absolute temperatures. 

The entropy of vaporization is obviously shown for each 
different temperature (or pressure) by the distance between 
the water curve and the saturation curve, since the former 
is distant from the vertical axis by an amount equal to fa, 
while the latter is distant an amount equal to fa -{-fa. 

Superheating lines are drawn by picking from the steam 
tables the values of entropy above 32° F. for steam super- 
heated to different temperatures at one particular pressure 
and plotting against the proper temperatures. There will 
be as many superheating lines on the diagram as one chooses 
pressures for which to plot them. Only one is shown 
in the figure. 

One very useful property of this diagram follows from the 
fact that points on its surface indicate the condition of the 
material. For instance, if the temperature-entropy, or 
T—4>, values of the material at a given condition should 
plot to the left of the liquid line, the material must be in the 
liquid condition; if they plot between the liquid line and the 
saturation curve, the material must be a mixture of liquid 
and saturated vapor; if they plot on the saturation curve, 
the material must be dry, saturated steam; and if they 
plot to the right of the saturation curve, the material must 
be superheated steam. This all follows directly from the 
definition of entropy above 32° F., as plotted in these dia- 



64 STEAM POWER 

grams. The various regions, or fields, into which the dia- 
gram divides in this way are shown in Fig. 23 (a). 

Another very useful property of this diagram follows 
from the fact that area represents heat just as area on a 
pressure-volume diagram was found to represent work. 
Thus the area under the line ab, for instance, represents 
the heat required to raise the temperature of one pound 
of water from 32° F. to the temperature at b. Similarly 
the area under the line be represents the heat required to 
change a pound of water at the temperature at b to a pound 
of dry, saturated steam at the same temperature. The 
heat required to superheat this pound of saturated steam 
at constant pressure up to the temperature shown at d 
is similarly represented by the area under the line cd. 

In this connection, it should be noted that this diagram 
is plotted above absolute zero of temperature just as the 
pressure-volume diagram is plotted above absolute zero of 
pressure. The areas in question therefore extend down 
to the absolute zero of temperature. In order to indicate 
this in Fig. 23 (b), a large part of the chart is supposed to 
have been broken out, so that the lower end of the diagram 
could be moved up into view. In Fig. 23 (a), the bottom 
of the diagram is drawn a few degrees below 32° F. and 
this is indicated by putting T>0 opposite the horizontal 
axis. 

The various areas hatched in Fig. 23 (b) indicate the 
various quantities of heat previously discussed. It should 
be understood that the areas represent the heat quantities 
only for the particular pressure which corresponds to the 
temperature indicated by T v . For a higher pressure, the 
line be would be higher and the areas proportionately 
larger; for a lower pressure the line be would be lower and 
the areas smaller. 



ENTROPY DIAGRAM 



65 



327.8 F. 



281°P. ' 



ILLUSTRATIVE PROBLEM 

Starting with liquid at a temperature T x corresponding to the 
temperature of vaporization at a pressure of 50 lbs. per square 
inch absolute, assume the liquid raised to the temperature of 
vaporization at a pressure of 100 lbs. per square inch absolute 
and then completely vaporized. Determine the various changes 
of entropy and indicate them on a TV-chart. 

The steam tables give entropy of the liquid, <fr, as equal to 
0.4113 for water about to vaporize under 50 lbs. per sq. in. 
absolute, and 0.4743 for water about to vaporize under a pres- 
sure of 100 lbs. per sq. in. absolute. The difference, that is, 
0.4743-0.4113=0.0630, must be 
the entropjr change experienced 
by the liquid when its tempera- 
ture is raised from the lower to 
the higher value. These values 
fare shown in Fig. 24. 

The steam tables give entropy 
of vaporization, </>„, at 100 lbs. per 
square inch absolute as 1.1277. 
Adding this to the entropy above 
32° F. of the liquid at vaporiza- 
tion temperature under 100 lbs. 
pressure gives 0.4743+1.1277 = 
1.602 as the entropy above 32° 
of dry, saturated steam at 100 

lbs. per square inch absolute. These values are all indicated in 
their proper position in Fig. 24. 

The total change of entropy experienced by the material in 
changing from water at the temperature of vaporization under 
50 lbs. pressure to dry, saturated steam at 100 lbs. pressure is 
obviously equal to 0.0630+1.1277 = 1.1907. 

43. Quality from T^-chart. The entropy change ex- 
perienced by steam in the process of vaporization is directly 
proportional to the addition of heat. Thus, when half 
the latent heat has been added to one pound of material, 
the entropy change is §<£». In general, if a fraction x of 
the latent heat has been added, the entropy change has 
been xfa during the process. Therefore, if the temperature 
entropy condition of a pound of material should plot at a 




Entropy (0) 




Fig. 24. 



66 



STEAM POWER 



point such as c in Fig. 25, it follows that the material is 

a mixture of water and steam and that a fraction of the 

be 
pound equal to — - is steam, the rest being water. But, 



bd 



be 



by definition, the fraction — is x, 



the quality of the 



material. 

The temperature-entropy chart is very useful when used 
in connection with this property of showing quality. Thus, 
in Fig. 25, the area under be, down to absolute zero tem- 
perature, represents the fraction of the latent heat of 




Entropy 



Fig. 25. — Quality from Temperature- Fig. 26. — Constant Quality 
Entropy Chart. Curves. 



vaporization per pound which must be added to give a 
pound the quality x. 

For convenience in use, constant quality lines are 
generally drawn on temperature-entropy charts. Such 
lines are shown in Fig. 26. Each line is obtained by plot- 
ting the temperature entropy conditions for a given quality 
at different pressures. For this purpose, 4> v and <f>i are 
taken from the steam tables for a given pressure. The 
numerical value of <£«, is then multiplied by the fraction re- 
presenting the chosen quality, say 0.9, and the product 
is added to <J>i, giving the total entropy above 32° F. for 
quality 0.9 at the particular pressure chosen. The same 



ENTROPY DIAGRAM 



67 







(\1 ) a-m^aodtnojt o;npsqv 



68 STEAM POWER \ 

process is repeated with the same value of the quality, 
but with different pressures, until enough points have been 
secured to make it possible to draw a smooth line through 
them. 

44. Volume from T<£-chart. Since quality changes at 
any given temperature, or pressure, are accompanied by 
volume changes, it is possible to find a series of values for 
the quality of a pound of wet steam which will make that 
pound occupy the same volume at different temperatures. 
Having found the quality which will be necessary at a num- 
ber of different temperatures, the total entropy above 32° 
F. can be found for each case and these values can then be 
plotted on the T</>-chart. Connecting the points so obtained 
would give what is known as a Constant Volume Line. 

Several of these constant volume lines are shown in their 
correct positions in Fig. 27. It will be observed that, 
for each volume, the quality must increase as temperature 
(and pressure) increases in order to maintain a constant 
value for the volume occupied by one pound of mixture. 

45. Heat from T^-chart. Equations for obtaining the 
total heat above 32° F. for wet and for superheated steam 
were given in an earlier chapter. By means of these 
equations, it is possible to find a succession of values for 
quality and superheat which will give a pound of material 
any chosen heat content at different pressures. If the 
corresponding values of temperature and entropy are found 
and plotted, what is known as a Constant Heat Line results. 
Several of these lines are shown in Fig. 27. 

46. The Complete T^-chart for Steam. A very com- 
plete, graphical representation of the properties of water 
and steam can be procured by combining in one diagram 
all of the lines discussed in preceding paragraphs. Such 
a diagram is generally spoken of as the T<t>-diagram or the 
T<j>-chart for steam. An example of such a diagram is 
given in Fig. 28. 

This chart is very useful, as it enables one to solve by 



•jqrjj; - o-irqiuaclmaj; 





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TEMPERATURE-ENTROPY DIAGRAM 

TO ACCOMPANY 

STEAM POWER 

C.F. HiRSHFELD AND T.C. ULBRICHT 


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70 STEAM POWER 

inspection many of the most difficult problems which arise 
in the theory and practice of using steam. As an example, 
assume that it is desirable to know what will happen if 
water at the temperature of vaporization corresponding 
to about 24 lbs. per square inch absolute has its volume 
increased indefinitely at constant temperature. The initial 
condition of the water would be shown on the water curve 
of Fig. 28 at the point at which the 700° absolute temperature 
line crosses it. Increase of volume at constant temperature 
would be indicated by a horizontal line running to the right 
from this point. Obviously, vaporization will occur at 
constant pressure (because the temperature is constant) 
and the quality will change from zero to unity at which the 
saturation curve will have been reached. Further increase 
of volume can result only in the production of superheated 
steam, since the line representing the process will run 
out into the superheated steam field. It is also interesting 
to note that the pressure on the material will have to be 
decreased as the volume increases in the superheated steam 
region, as is evidenced by the fact that the horizontal line 
representing the assumed process cuts lower and lower 
pressure lines as it is extended to the right in the super- 
heated field. 

Note also that the intersections of this horizontal line 
with constant volume and constant heat lines afford the 
means of determining volume and heat above 32° F. at 
different stages of the assumed process. 

PROBLEMS 

1. Determine from the steam tables the change of entropy 
experienced by one pound of water when its temperature is raised 
from 32° F. to the temperature of vaporization under a pressure 
of 100 lbs. per square inch absolute. 

2. Determine from the steam tables the entropy change experi- 
enced by one pound of water when its temperature is raised from 
32° F. to the temperature of vaporization under a pressure of 
150 lbs. per square inch absolute. 



ENTROPY DIAGRAM 71 

3. Determine the entropy change experienced by one pound of 
water when its temperature is raised from the temperature of 
vaporization corresponding to 100 lbs. per square inch to that 
corresponding to 150 lbs. per square inch by subtracting the value 
found in Prob. 1 from that found in Prob. 2. 

4. Determine the change of entropy experienced by one pound 
of material completely vaporizing at a temperature of 327.8° F. 

5. Plot a T</>-chart for one pound of water. Start by plotting 
entropy of the liquid for various temperatures; then plot entropy 
of saturated steam (above 32° F.) ; finally draw water line, satura- 
tion line, and several lines showing change of entropy during vapor- 
ization. 

6. Determine from a T^-chart the quality which would be 
attained by one pound of steam if it experienced a change which 
carried it from the condition of dry saturated steam at 150 lbs. 
per square inch absolute to a pressure of 25 lbs. per square inch 
absolute by a process which would plot as a vertical line on the 
7>-chart. 

7. Assume a pound of mixed water and steam to have a qual- 
ity of 80% at a pressure of 200 lbs. per square inch absolute. 
Determine from the T^-chart the heat above 32° per pound of 
mixture and the volume occupied by the mixture. Determine 
also the quality attained if the pressure of the material drops to 
20 lbs. per square inch absolute at constant entropy. How does 
the heat above 32° F. change during such a process? 

8. Assume a pound of mixture as in Prob. 7, but with a 
quality of 30% at a pressure of 200 lbs. Find all quantities called 
for in that problem. 

9. Assume a pound of material as in Probs. 7 and 8 above, 
but superheated 200° at a pressure of 200 lbs. per square inch 
absolute. Determine all quantities called for in Prob. 7. 

10. Choose a point on the TV-chart at which a constant volume 
line intersects the saturation curve. Determine the change of 
quality, entropy and heat above 32° F., if the material drops 
to half pressure at constant volume. 



CHAPTER VI 

TEMPERATURE ENTROPY DIAGRAMS OF STEAM 
CYCLES 




47. Complete Expansion Cycle. This cycle was con- 
sidered in Chapter IV and the PF-diagram was given there 
as Fig. 21. The diagram of this cycle drawn to T^-co- 
ordinates is shown in Fig. 29. The same letters are used 

to represent corresponding points 
in the two diagrams. 

The entropy change during 
the heating of the liquid is 
shown by the part of the liquid 
line between d and a, and the 
heat supplied during that process 
is represented by the area below 
the line da, measuring clear 
down to the absolute zero of 
temperature. 

The entropy change during 
vaporization is represented by the line ab and the heat 
supplied during the process is shown by the total area 
under that line. 

The adiabatic expansion of the steam is represented by 
the line be, such an adiabatic change fortunately being a con- 
stant entropy process and therefore easily drawn in this 
diagram. Obviously no heat is received or removed dur- 
ing this process, as there is no area under the line be. 

The entropy change during condensation is represented 
by the line cd and the heat rejected by the working sub- 
stance during this process is represented by the area under 

that line. 

72 



Fig. 29.— 7Vdiagram, Com- 
plete Expansion Cycle. 



TEMPERATURE ENTROPY DIAGRAMS 73 

48. Area of Cycle Representative of Work. It will 
be remembered that area under a line in the PF-diagram 
represents work in foot-pounds. That diagram, however, 
gives no indication of heat received or rejected and it is 
not possible to obtain any direct idea of efficiency from it. 
In this respect, the T^-diagram is much better. Area under 
the lines da and ab in Fig. 29 represents heat supplied 
the working substance. Area under the line cd represents 
heat rejected by the working substance. The difference 
between these two, or the area enclosed within the lines 
of the cycle, must therefore represent the heat converted 
into mechanical energy per cycle. 

This diagram therefore shows directly by areas the 
heat supplied, the heat rejected, and the heat converted 
into mechanical energy. Further, the ratio of the area 
representing heat converted into work, and the area repre- 
senting heat supplied must be the efficiency of the cycle. 

Remembering also that if the lines of the cycle are drawn 
upon a T^-chart such as that given in Fig. 28, all volume 
changes, heat contents and qualities at different points 
are shown without further work, it becomes evident that 
this form of representation is decidedly convenient and far 
superior to the pressure volume method. 

49. Modifications for Wet and Superheated Steam. 
The complete expansion cycle is supposed to represent an 
idealization of what happens in a real prime mover. In real 
cases, however, the steam may arrive at the prime mover wet 
or superheated and it is desirable to investigate the method 
of representing such conditions as well as their effects. 

Wet steam corresponds to incomplete vaporization, 
i.e., a quality less than unity at the upper right-hand corner 
of the cycle. This might be shown for a given case by the 
location of the point 6' in Fig. 29. The cycle would then 
be ab'-c'd and a smaller amount of work would be obtained 
per pound of working substance as evidenced by the smaller 
area enclosed within the lines of the cycle. 



74 



STEAM POWER 



In the case of superheated steam, superheating occurs 
at constant pressure after vaporization is complete. This 
would be shown by the location of the upper right-hand 
corner of the cycle at some point b" on the constant pressure 
line which extends out from b. The cycle is now represented 
by abb"c"d and evidently has a different shape than it 
had in the preceding cases. Obviously the area enclosed 
within the lines of the cycle is greater than it was before and 
therefore more mechanical energy is obtained per pound 
of steam. 

50. Incomplete Expansion Cycle. The only difference 
between the incomplete and complete expansion cycles is 




Fig. 30. — T<£-diagram, Incom- 
plete Expansion Cycle. 



Fig. 31. — 'TV-diagram, Cycle 
Without Adiabatic Expansion. 



the termination of the expansion in the former by means of 
a constant volume line. This is shown to T^-coordinates 
in Fig. 30 in which the incomplete expansion cycle is drawn 
in heavy lines over the one in which expansion continues 
to the back pressure. 

i The constant volume line is seen to cut off a corner, 
thus reducing the area representing heat converted into 
work. The heat supplied in each case is measured by the 
area under the lines ea and ab. The efficiency of the cycle 
with incomplete expansion can therefore be seen to be less 
than that of the other cycle by simple inspection of the 
diagram. 

If the adiabatic expansion is terminated at a higher 



TEMPERATURE ENTROPY DIAGRAMS 75 

pressure, as by the constant volume line c"d" in Fig. 30 r 
still more of the work area is lost, but the same quantity 
of heat is supplied, and therefore the efficiency is still lower 
than when the expansion terminated at c. Obviously 
as the point at which the adiabatic expansion is terminated 
moves nearer and nearer to b as shown in Fig. 31, the cycle 
becomes less and less efficient. If the constant volume 
line starts at b, there is no adiabatic expansion and the 
cycle becomes that previously considered as having a rec- 
tangular shape in the PF-diagram. This cycle has the shape 
indicated by abed in the T^-diagram of Fig. 31. Obviously 
it is least efficient of all as was previously shown by other 
means. 

51. Effect of Temperature Range on Efficiency. It has 
already been stated (see p. 26) that heat engines receive 
heat at a high temperature, convert some of it into me- 
chanical form and discharge the remainder at a lower tem- 
perature. Inspection of the T^-diagram shows this very 
clearly, and, remembering that the area of the cycle measures 
the heat converted, these diagrams also show how raising 
the upper temperature (or pressure) or lowering the lower 
temperature (or pressure) will increase the efficiency. It 
can be seen readily that lowering the lower temperature 
will, however, be more effective in increasing the efficiency 
than raising the upper temperature. 

PROBLEMS 

1. Draw a complete expansion cycle to TV-coordinates for the 
following conditions (using ^-diagram for steam to get values); 
weight of working substance, 1 lb.; initial pressure, 125 lbs. 
absolute; quality at beginning of adiabatic expansion, 100% 
back pressure, 10 lbs. absolute. 

2. Determine the following values for cycle drawn in Prob. 1 : 

(a) Entropy of liquid at beginning of vaporization; 

(b) Entropy at beginning of adiabatic expansion; 

(c) Quality at end of adiabatic expansion; 

(d) Volume at end of adiabatic expansion- 

(e) Entropy at end of condensation. 



76 STEAM POWEP 

3. Show by measuring the area on T ^-diagrams, the increase of 
efficiency resulting from the use of an initial pressure of 175 lbs. 
absolute and from the use of a terminal pressure of 2 lbs. absolute 
in place of the values given in Prob. 1. 

4. Compare the efficiency of a complete expansion cycle with 
conditions as in Prob. 1 with a complete expansion cycle with 
same pressures but with a temperature of 500° F. at the beginning 
of the adiabatic expansion. 

6. Draw an incomplete expansion cycle to T^-coordinates 
for the same pressures as in Prob. 1, but with adiabatic expan- 
sion ending at a pressure of 15 lbs. absolute. 

6. Compare work and efficiency of the two cycles of Probs. 
1 and 5 above. 

7. Draw a cycle without expansion for the conditions of 
Prob. 1 to TV-coordinates and compare the work area with that 
obtained in Probs. 1 and 5. 






CHAPTER VII 



THE REAL STEAM ENGINE 

52. Operation of Real Engine. In previous chapters 
the ideal steam engine was considered and several cycles 
upon which it might be operated were discussed. Real 
engines are built to operate on the same cycles, but because 
of certain practical considerations they only imperfectly 
approximate the ideal performance. 

Real engines must be built of iron and steel for practical 
reasons and these metals absorb, conduct and radiate heat 
so that certain heat interchanges between the working 
substance and engine and certain heat losses occur in 
practical operation. These were eliminated in the ideal 
case by simply assuming ideal materials not possessed 
of the characteristics of real metals. 

It is also practically impossible to generate steam 
in the cylinder of a real engine as was assumed to be done 
in the ideal case. Heat is practically obtained by the com- 
bustion of fuels, and the higher the temperature attained 
the better can the liberated heat be utilized in the genera- 
tion of steam. To subject the cylinder to such high tem- 
peratures and to control the heating and cooling as neces- 
sary to produce a number of cycles in rapid succession would 
lead to rapid wear and great practical difficulties. It has 
been found best to generate the steam in a boiler which is 
properly equipped for that purpose and then to transmit 
it with its contained heat to the engine, which is constructed 
in such a way as to utilize that heat to the best advantage. 
If the steam is to be condensed, as assumed in the ideal 
case?, it has also been found best to remove it from the 

77 . 



78 



STEAM POWER 



cylinder and to condense it in a separate piece of appartiTus 
properly constructed for that purpose. 




'co 

o 

o 

o 

h 

o 

u 

a 



H 



The entire arrangement which results from these prac- 
tical modifications in the case of a non-condensing engine 




THE REAL STEAM ENGINE 79 

is shown in Fig. 32. Steam is generated within the boiler 
at some constant pressure Pi and at the proper instant the 
admission valve at one end of the cylinder is opened, allow- 
ing steam to flow in and drive the piston outward. If 
there were no losses, this would be represented by some 
such line as ab at a height Pi on the 
PF-diagram of Fig. 33. Closing of 
the valve after the piston had moved 
part way out would cut off the 
further flow of steam, and, with con- 
tinued motion of the piston, the steam 

within the cylinder would expand. ^ QQ 

r ig. oo. 

If no heat interchanges occurred, 

this expansion be would be adiabatic as in the real case. 

It will be observed that the two lines on the PF-diagram 
thus far produced represent equally well the corresponding 
two lines of the complete or incomplete-expansion cycles. 
The heat supplied in the boiler is the same as that supplied 
in the cylinder under the ideal conditions originally assumed, 
and the work under the line ab is equal to the external 
work done during vaporization just as in the ideal case. 
If difficulty is experienced in connection with the statement 
regarding external work, it is only necessary to picture the 
process in this way: Assume that each pound of steam 
formed in the boiler does the external work equivalent to 
APu by pushing the pound previously generated ahead of 
it as a piston, and that this motion communicated along 
the pipe from layer to layer results in pushing an equivalent 
weight (aud volume) into the cylinder against the resist- 
ance offered to the piston's motion. 

When the piston arrives at the end of its stroke at the 
point c, the opening of the " exhaust valve," connecting 
the interior of the cylinder with the space in which the 
pressure P2 lower than P c is maintained, will permit some 
of the steam to blow out of the cylinder with the piston 
standing stationary at the end of its stroke. This would 



80 STEAM POWER 

give a constant volume change equivalent to the correspond- 
ing line in the incomplete-expansion cycle. 

The return of the piston from d to e, with the exhaust 
valve still open, would force the remainder of the steam 
out of the cylinder and into the space in which the pres- 
sure P2 is maintained. The result, so far as the diagram 
is concerned, is obviously the same as in the ideal case, and 
if the steam were condensed within a proper vessel into 
which it exhausted (instead of being exhausted to atmos- 
phere), the result would also be the same so far as the shape 
of the diagram is concerned. The pressure P2 might, how- 
ever, be maintained at a lower value, thus giving a greater 
temperature range. 

The pressure rise ea within the cylinder would result 
directly from the opening of the admission valve and the 
admission of steam for the next cycle. But, if the working 
substance is to be returned to starting conditions as was 
dene in the ideal case, its pressure must also be raised to 
Pi and its temperature to a corresponding value. The 
pressure is raised in the case of condensing operation by 
means of the boiler feed pump, which picks up the condensed 
steam (condensate) and forces it into the boiler. The 
temperature of the working substance is raised by passing 
it through feed-water heaters or by heat absorbed directly 
from the heated water in the boiler. 

When operating non-condensing the working substance 
exhausted during the last part of each cycle is really thrown 
away by allowing it to mix with the atmosphere, but an 
equivalent quantity of water is fed to the boiler by the 
boiler feed pump and takes the place of the material lost 
by exhaust to atmosphere. This method of operating 
does not approximate the ideal as closely as does the con- 
densing method, but the discrepancy is not very great. 

53. Losses in Real Installations. The diagram given 
in Fig. 33 was obtained by assuming the absence of certain 
practical losses and is considerably modified when real 




THE REAL STEAM ENGINE 81 

apparatus is used. Thus the real engine, as shown in con- 
nection with Fig. 9, has clearance and operates with com- 
pression so that the clearance is filled with steam at a pres- 
sure indicated by the point a' in Fig. 34 when the admission 
valve opens. 

There is also always some drop of pressure along the 
steam pipe so that the pressure 

. . i • i . -t i±t ,^— Adiatatic Expansion 

at the engine is lower than at the p, — — ^\ of mixm™ at point 

of Cut-off. 

,Adiaba 

. X all Ma 

valve can never be made to give a ^'Mo 

such a large opening into the 

cylinder that there is not a 

measurable drop of pressure in FlG . 34 ._ Theoretical and Real 

flowing through it. As a result indicator Diagrams. 

of these actions the highest 

pressure attained within the cylinder as indicated at point 

a in Fig. 34 is always lower than the boiler pressure P\. 

As the piston of a real engine moves out it acquires 
a higher and higher velocity until it reaches a point near 
mid-stroke. The entering steam therefore must flow 
through the valve with increasing velocity if it is to follow 
up the piston and fill the cylinder, but this usually neces- 
sitates greater pressure drops as the piston moves out, 
so that the admission line generally slopes downward instead 
of being horizontal. There is also another phenomenon 
which causes this line to slope. The metal of cylinder, 
cylinder head and piston is in contact with comparatively 
low-temperature steam during the latter part of each cycle, 
and therefore acquires a lower temperature than that 
of the steam about to enter. Therefore, when the high- 
pressure steam enters the cylinder it gives up heat to 
the walls at a comparatively rapid rate, and, if initially 
dry saturated, this results in a great deal of condensation. 
Such condensation is called initial condensation. 

As the steam condenses after flowing into the cylinder 
and forms water occupying a negligibly small volume, 



82 feTEAM POWER 

it follows that steam must flow into the cylinder at a pro- 
portionately greater rate in order to fill the space vacated 
by the piston. But this results in an increased pressure 
drop and therefore would give a sloping admission line. 

When the piston has finally been driven out as far as 
desirable by the action of high-pressure steam, the admis- 
sion valve is closed, that is, cut-off occurs. This valve can 
not be closed suddenly; the closure is more or less gradual 
in all cases. As the opening becomes smaller it becomes 
increasingly more difficult for the steam to flow through 
and into the cylinder so that the pressure continues to drop 
at an increasing rate until the valve is finally closed. This 
gives the rounded cut-off shown at the point b. 

The loss of pressure during admission is generally 
said to be due to throttling or wire drawing, these terms 
being intended to convey the idea that the steam has to 
squeeze its way through the inlet openings with correspond- 
ing loss of pressure. 

When the cut-off has finally been completed, it leaves 
the end of the cylinder filled with a mixture of steam and 
water at steam temperature, and this mixture then expands 
as shown by the line be. At the beginning of the expansion 
the steam generally has a higher temperature than that of 
the surrounding walls and it therefore continues to give 
heat to those walls. Were the expansion adiabatic it 
would follow the dot-dash line in the figure, but, as the 
steam must not only convert heat into work, but must also 
supply heat to the walls, it condenses more rapidly than 
in the ideal case and its pressure and volume changes follow 
some such law as that indicated by the upper part of the 
curve be. 

As expansion continues, the pressure and temperature 
of the steam drop until some point is reached at which 
the temperature has become equal to that of the walls. 
Further expansion with drop of pressure and temperature 
results in reducing the temperature of the steam below 



THE REAL STEAM ENGINE 83 

that of the walls, and then the direction of heat transfer 
is reversed, the hot walls giving heat to the cooler steam 
at an increasingly rapid rate. This heat causes re-evapora- 
tion of some of the water formed before and thus tends to 
increase the volume occupied by the material in the cylinder, 
with the result that the lower part of the expansion curve 
be approaches and generally crosses the curve which would 
have been attained by adiabatic expansion in non-conduct- 
ing apparatus. 

In many real engines the re-evaporation is so great that 
the steam is entirely dried and sometimes superheated 
before the exhaust valve opens. 

The exhaust valves of steam engines are always opened 
before the piston reaches the end of the stroke, as it is found 
necessary to do this if excessive losses are not to occur 
due to the difficulty of forcing the large volume of low- 
pressure steam through the exhaust passages. When opened 
early enough, the steam flows out in such quantity before 
the end of the stroke that the " back pressure " during the 
return or exhaust stroke is only a pound or two above that 
of the space into which the engine is exhausting. 

During all of the exhaust period, the steam is probably 
at a lower temperature than the walls to which it is exposed 
and re-evaporation probably continues in most cases until 
the closure of the exhaust valve. It seems probable that 
the steam retained in the cylinder after the closure of the 
exhaust valve is approximately dry, but little is really 
known regarding the quality of the clearance steam. 

The rise of pressure during compression has two bene- 
ficial effects: It helps to bring the moving parts to rest grad- 
ually, and it raises the temperature of the clearance steam 
and of the walls of the clearance space to values nearer 
that of the entering steam. 

Remembering that area on a PT-diagram represents 
work, it is easily seen that throttling losses and rounding 
of corners due to slow valve action (which cause a loss of 



84 STEAM POWER 

diagram area) result in a loss of work. The fact that conden- 
sation also causes a great loss is easily shown. A given 
quantity of steam entering the engine with its supply of 
heat can, in the ideal case, do a certain amount of work at 
the expense of that heat. In the real case part of the heat 
is stored in the walls during the early part of the cycle, 
so that it is not available for the doing of work and is removed 
from the walls and carried out into the exhaust as unutilized 
heat during the later part of the cycle. The phenomenon 
can be pictured by imagining the steam as dropping some 
of its heat into a pocket in the walls of the cylinder when 
entering the engine and then picking it up again and carrying 
it out when leaving, so that the next charge of steam will 
have to fill the pocket again. 

The net result of condensation and re-evaporation 
is the obtaining of less work from a given quantity of steam 
than should be obtained, or the use of more steam than 
theoretically necessary for a given quantity of work. This 
effect is shown graphically by the two adiabatic expansion 
lines of Fig. 34. 

The initial condensation in real engines which are sup- 
plied with saturated steam generally amounts to from 20 to 
50 per cent of all the steam supplied, so that it is evident 
that anything which will prevent part or all of this loss should 
do much to improve the steam consumption of engines. 
This subject will be discussed in more detail in later para- 
graphs and various methods of decreasing losses from this 
source will be considered. 

54. Clearance. The term clearance is used in a two- 
fold sense; (a) to refer to mechanical clearance or the 
linear distance between the two nearest points of cylinder 
head and piston face when the piston is at the end of its 
stroke, and (b) to refer to volumetric clearance or the 
volume enclosed between the face of the valve, the cylinder 
head and the face "of the piston when the latter is at the 
end of its stroke. 



THE REAL STEAM ENGINE 



85 



Ste-.m Port Steam Chest Valv 







The former is generally given in inches and varies 
from a very small fraction of an inch in the best engines 
to an inch or more in cheap and in poorly designed engines. 
It is indicated by a in Fig. 35. 

The volumetric clearance is expressed as a percentage 
of the piston displacement or 
volume swept through by the 
piston. It varies from 2 per 
cent or less in the best engines 
to as high as 15 per cent in 
the cheaper and less economical 
models. It is made up of the 
parts designated by c in Fig. 35. 

55. Cushion Steam and Cyl- 
inder Feed. It is customary 
to imagine the steam operating 
within an engine cylinder to FlG " 35.-Mechanical and Volu- 

. . » . . , , 7 . metric Clearances, 

consist ol two parts, the cushion 

steam and the cylinder feed. The former is that part of 
the total which is contained in the clearance space before 
the admission valve opens and serves to cushion the 
reciprocating parts of the engine. The cylinder feed 
is the steam which enters through the valve for each 
cycle. 

If the same cycle is produced time after time so that all 
temperature effects are repeated at regular intervals and 
so that all events occur at the same points in successive 
cycles, the quantity of steam retained in the clearance 
volume will be the same for successive cycles. It is 
impossible to measure the quantity of this steam directly 
and indirect methods are therefore adopted for that 
purpose. 

It is often assumed that the steam is dry and satu- 
rated when compression begins, as at the point e in Fig. 
34. With this assumption, the weight of cushion steam 
can be determined by dividing the volume occupied, that is, 



86 STEAM POWER 

V e , by the volume occupied by one pound of dry saturated 
steam at the same pressure. Thus, 

Ve 

Cushion steam = = ; — — — lbs. . (29) 

bp.vol. at pressure P e 

The weight of cylinder feed can be very accurately 
determined by condensing and weighing the steam leaving 
the engine in a given time and dividing by the number of 
cycles performed during the same period. It can also be 
determined by metering the steam entering the engine 
or by measuring the water fed to a boiler supplying only the 
engine in question. An approximate determination of the 
quantity of the cylinder feed can also be made directly from 
an indicator diagram by determining what is known as the 
diagram water rate. This will be considered in detail at 
a later point. 

When cushion-steam and cylinder-feed have both been 
determined, the weight of steam contained in the cylinder 
between cut-off and release can be found by adding the two 
quantities. Thus, 

W = W f +W K , (30) 

in which 

W = total weight of steam expanding in cylinder per cycle ; 
Wf= weight of cylinder feed per cycle; and 
Wk = weight of cushion steam per cycle. 

The volume which the mixture would occupy if dry 
and saturated at any given pressure can be determined by 
multiplying W the total weight by the specific volume for 
that particular pressure. 

56. Determination of Initial Condensation. The loss 
due to initial condensation is so important that it is cus- 
tomary to determine the amount of this loss when studying 
engines. This can be done with fair accuracy by means 
of the indicator diagram. 



THE REAL STEAM ENGINE 



87 



To make such a study it is necessary to know the total 
weight of material in the engine cylinder at the point of 
cut-off. This weight may be determined by any of the 
methods just given. With the weight known, the volumes 
which this material should occupy at different pressures 
if dry and saturated can be determined by multiplying by 
the specific volumes at the various pressures. Plotting 
these points on a PF-diagram and connecting them will 
give a saturation curve for the material in the cylinder such 
as the curve shown in Fig. 36. 

By drawing this curve on the indicator diagram ob- 
tained from the engine and then comparing distances 
such as ab and ac as explained in section 26 of Chapter III 





Fig. 36. 



Fig. 37. 



the quality of the steam within the cylinder at all pressures 
between cut-off and release can be determined. The weight 
of initial condensation (up to the point of cut-off) must 
be the total weight of water shown as existing within the 
cylinder at that point minus any water brought in by the 
steam if it was not dry when entering the engine. 

Should the saturation curve cross the real expansion 
curve, as shown in Fig. 37, it indicates that the steam oc- 
cupies volumes greater than the specific volumes toward 
the end of the expansion; the steam within the cylinder 
must therefore be superheated during this part of the 
cycle. 

Many formulas have been devised for giving the quan- 
tity of initial condensation. They are all based upon the 
results of experiment and generally only give reliable 



88 STEAM POWER 

values for cases similar to those used in developing them. 
One formula of this sort which has been very widely tested 
and been found to give reliable results within its field 
of applicability is that devised by Robert C. H. Heck and 
explained in his books on the steam engine. The formula is 



~£=JA • (31) 



m 

in which 

ra = the fraction representing initial condensation; for 
ordinary cases it is the fraction of the cylinder 
feed which is condensed during admission, but when 
compression is very high and when great weights 
of steam are retained in the clearance it is the frac- 
tion of all the material within the cylinder which 
exists in liquid form at the time of cut-off; 
c = a coefficient, which varies between 0.25 and 0.30 with 
type of engine. May be assumed at 0.27 for average 
work; 

N = engine speed in revolutions per minute (R.P.M.); 
s = a constant for any engine, equal to nominal surface in 
square feet divided by nominal volume in cubic 
feet. The nominal surface is the area of the inner 
walls and the ends of a cylinder with diameter equal 
to the internal diameter of the cylinder and with 
a length equal to the stroke of the engine. The 
nominal volume is the cubic contents of such a 
cylinder; 

s = — ( 2— +4 j in which D and *S represent diameter and 

stroke of engine in inches; 
6 = a temperature function obtained from Table II as 

there indicated; 
p = the absolute pressure in cylinder in pounds per square 

inch just after completion of cut-off; 



THE EEAL STEAM ENGINE 



89 



e = cut-off ratio, that is, ratio of cylinder volume opened 
up by time cut-off has just been completed to the 
total piston displacement. 

TABLE II 

For Finding Values of 6 for Use in Heck Formula 



9=ki 


— kt when k\ 


and k 


2 are 


jhosen 


"rora table for 


highest and lowest pressures 






existing in cylinder 




V 


k 


V 


k 


V 


k 


P 


* ! 


V 


k 


V 


k 


1 


175 


15 


210 


50 


269.5 


90 


321.5 


160 


389 


230 


441 


2 


179 


20 


220 


55 


277 


100 


332.5 


170 


397 


240 


447.5 


3 


183 


25 


229 


60 


284 


110 


343 


180 


405 


250 


454 


4 


186 


30 


238 


65 


291 


120 


353 


190 


413 


260 


460.5 


6 


191 


35 


246 


70 


297.5 


130 


362 . 5 


200 


420 


270 


467 


8 


196 


40 


254 


75 


304 


140 


371.5 


210 


427 


280 


473 


10 


200 


45 


262 


80 


310 


150 


380.5 


220 


434 


290 


479 



57. Methods of Decreasing Cylinder Condensation. 

Before discussing methods of decreasing the loss due to cylin- 
der condensation it will be well to consider what things may 
be expected to determine the extent of such loss. The 
condensation is due directly to the transfer of heat from 
one body to another at lower temperature, and anything 
which tends to increase the total amount of heat thus trans- 
ferred will increase the total condensation. 

It is therefore evident that the ratio of steam condensed 
to steam supplied will be greatest when: 

(a) The time of contact is greatest; 

(b) The ratio of surface exposed to volume enclosed is 
greatest, and 

(c) The temperature difference is greatest. 

The time of contact can be controlled to a certain 
extent by controlling the speed of the engine and, with 
other things equal, the higher the speed the lower should 
be the condensation. 

The ratio of surface exposed to steam to the volume 
occupied by steam has a great influence on the amount of 



90 STEAM POWEE 

condensation which occurs. The surface of the clearance 
space, including the interior surfaces of all ports or passages 
leading to the valves, seems to have the greatest influence, 
and the clearance space which is most nearly a short cylinder 
without connected passages may be expected to give the 
least initial condensation. 

The size of the engine is also important in this connec- 
tion. The area exposed does not increase as rapidly as 
does the volume inclosed when the diameter of a cylinder is 
increased, and therefore large cylinders give smaller ratio 
of surface to volume and therefore a smaller percentage 
of steam condensed. Large engines thus have a decided 
advantage over small engines. 

The shape of the cylinder also has an effect. The 
longer the cylinder with respect to its diameter the more 
favorable its performance. 

The point at which cut-off occurs is also intimately 
connected with the condensation loss. In a given cylinder 
with a given clearance the total condensation within the 
clearance space may be assumed practically constant if 
speed and temperature remain about the same. But if 
the cut-off is made later larger quantities of steam are 
admitted per stroke, and hence the fraction of the total 
cylinder feed which is condensed decreases. 

The temperature differences depend on upper and 
lower pressures, that is, on the pressure range. The inner 
surfaces of the walls follow as rapidly as possible the tem- 
perature changes of the steam within them. Thus their 
average temperature is somewhere between the upper and 
lower temperatures of the steam. If now, with a given 
upper steam pressure and therefore temperature, the lower 
pressure be reduced, the average wall temperature also will 
be reduced, and therefore the differences between the 
temperature of the entering steam and the average tem- 
perature of the walls will be increased with a resulting in- 
crease in condensation loss. 



THE REAL STEAM ENGINE 91 

The methods of decreasing this loss can now be con- 
sidered. They are given below under separate heads 
with brief explanation when necessary. 

(a) Clearance spaces should be properly designed so 
that the minimum surface is exposed. 

(6) The proportions of cylinder (diameter and stroke) 
and the speed of the engine should be so chosen that the 
condensation loss is reduced to a minimum. 

(c) The engine should be so proportioned that when 
delivering its rated power the cut-off occurs at such a point 
as to make the percentage of cylinder condensation the 
minimum consistent with other requirements. 

(d) The cylinder should be surrounded by spaces filled 
with air or by materials which are poor conductors of heat 
so as to decrease loss by radiation, because all heat lost in this 
way must be supplied by the condensation of steam within 
the cylinder. Such metallic parts as cannot be " lagged " 
in this way should be polished because polished surfaces 
radiate less heat than dull surfaces under like conditions. 

(e) The cylinder may be surrounded by a steam jacket, 
that is, a space filled with steam similar to that supplied 
the cylinder. The use of such a jacket sometimes results 
in a considerable saving and at other times in a great loss. 
The cylinder proportions, speed and pressure range seem 
to be the determining factors, and most long-stroke cylin- 
ders operating at low rotative speed and with small pressure 
ranges are jacketed. 

(/) The engine may be compounded, that is, the expan- 
sion of the steam may be made to occur in two or more 
cylinders taking steam in series. This results in decreas- 
ing the pressure range in each of the cylinders and effects 
a decided saving under proper conditions. Compounding 
will be considered in detail in a later chapter. 

(g) The engine may be supplied with superheated steam. 
If the steam is sufficiently superheated it can give up part 
or all of its superheat to heat the cylinder walls, and thus no 



92 



STEAM POWER 



condensation need occur. Heat interchanges between 
metal and superheated steam also appear to be less rapid 
than is the case when the steam contains water, so that a 
saving results from this source also. 

Tests made with saturated and with superheated steam 
indicate that from 7° to 10° of superheat are generally 
required to prevent 1 per cent, of initial condensation. 
Results differ greatly with the character of the engine, with 
its economy on saturated steam, with its valve gear, etc. 
Superheats of from 25° to 50° can generally be used with 
any well-designed engine, but higher temperatures usually 
require specially constructed engines. Superheats as high 
as 150° F. above saturation temperature are now quite 
common, and there seems to be a tendency to consider a 
value between 150° and 200° as the highest that is com- 
mercially advisable under ordinary conditions. 

58. Classification of Steam Engines. Steam engines, 
are classified on many different systems, the one used in 
any particular case being determined largely by circum- 
stances. The principal methods of classification are indi- 
cated in the following schedule : 



Classification of Steam Engines 
On the basis of rotative speed 



Low speed 
Medium speed 
High speed 

On the basis of ratio of stroke to diameter \ ». ± 

J J [Short stroke 

D-slide valve 



On the basis of 
valve gear 



Slide valve 



Balanced slide valve 
Multiported slide valve 
Piston valve 



~ .. , [Drop cut-off 
Corliss valve \ ~ ... . , 

[Positively operated 

Poppet valve 



THE REAL STEAM ENGINE 



93 



On the basis of position of longitudinal axis 



Vertical 
Inclined 
Horizontal 



On the basis of 
number of cyl- 
inders in which 
steam expands 



On the basis of cylinder arrangement 



On the basis of use 



Single expansion or 

Simple engine 

- r ,,. . f Compound expansion 

Multi-expansion „ . . 

. \ Triple expansion 

engine L~ , . 

I Quadruple expansion 

Single cylinder 
Tandem compound 
Cross compound 
Duplex 

Stationary engines 

Portable engines 

Locomotive engines 

Marine engines 

Hoisting engines 



59. Rotative Speeds and Piston Speeds. High-speed 
engines operate at a comparatively high rotative speed and 
are characterized by a short stroke in comparison with the 
diameter of the cylinder, the stroke generally being equal 
to, or less than, the diameter. The piston speed, by which 
is meant the feet travelled by the piston per minute, generally 
falls between 500 and 700. 

It is not considered advisible to allow piston speeds of 
stationary steam engines to exceed about 750 feet per 
minute for ordinary constructions and the great majority 
of engines give much lower values. The piston speed will 
obviously be given by the formula 



S = 2LN, 



(32) 



in which 



S = piston speed in feet per minute ; 
L = stroke in feet ; and 
N = revolutions per minute, 



94 



STEAM POWER 



and it is evident from this formula that as the rotative 
speed is increased the piston speed will increase unless 
the length of stroke is proportionately decreased. As 
a result, high-speed engines have short strokes in com- 



30 








\ 










































375 


•28 








\ 


\ 








































350 










\ 








































26 










\ 










H 


IGH 


SP 


EE 


D 


EN 


Gl 


NE 


S 












325 












\ 






































24 












\ 


£ 


~ 


































300 
















b. 


































22 
















V 


> 
































275 


















v6 
































1 

-3 20 

a 




































































































250 


a is 


















































225 pi 






















.«< 




























.3 16 

o 
u 






















5/ 
















































#/ 






























S00 g 200 
3 


14 


















4 
































700 | 175 
















^c" 


/ 






-p 


iftt* 


nj 


ipe 


>d 


















12 














J 


y 


■\: 


d^. 


s- 
























sed-Ft 












4 






































10 


















































on Sp 








^ 


c / 








































8 






< 


*/ 










































400 S loo 




& 


&>j 












































6 























































































































































8 10 12 14 16 18 20 22 
Diameter of Cylinder (Inches) 



24 26 28 



Fig. 38.— Proportions of High Speed Engines. 



parison with their cylinder diameters and slow-speed 
engines have long strokes. 

The characteristic relations between cylinder diameter 
and stroke, rotative speed and piston speeds of high-speed 
engines are given in Fig. 38. 

High-speed engines are generally fitted with some 



i 



, THE REAL STEAM ENGINE 



95 



2 S "§ 



R. P. M. . 

OS C5 CO CO ' 

Piston Speed (S) Ft. per Min. 















^ 
^ 


o 

8 


< 


g 


i 


i 










c 


I 




































































































\ 
























/ 


















\ 


\ 






















/ 




















\ 




















/ 






















\ 


v < 


1 
















/ 
























\ 
















/ 


























N 


\ 














/ 




























^ 












> 


/ 




























V 












/ 






























\\ 












/ 






























\ 


\ 








/ 
































\ 


\ 








/ 
































<5\ 

o 

o 




\ 




/ 


































*&\ 


\ 


k 


/ 


























UJ 








I* \ 




V 




























n 










p*\ 


V 


/' 




























z 












V 




-X 


























D 












A 




i s 


























UJ 








■ 


/ 




\ 




f\ 
























00 










/ 




\ 




* 


X 






















£ 








¥ 












*\ 






















_J 






















\ 


























/ 


7 














\ 


























/ 


















\ 
























/ 


















\ 












































\ 










































\ 


\ 








































^ 


\ 















































































































































































































































































































5 § 

.9 ° 



O "* ^ CO 

Stroke (L) in Inches' 



96 STEAM POWER 

form of balanced slide valve, and are controlled by what 
is called a shaft governor. They are very compact, having 
small weight and occupying small space in comparison with 
the power developed. 

Slow-speed engines are, in general, the most economical 
and are characterized by low rotative speed, long stroke, 
and elaborate valve gear. The weight per horse-power 
is high, and they generally occupy a great deal of space. 
The Corliss engine is the best known and most widely built 
engine of this type. 

The characteristic relations of cylinder diameter to 
stroke, rotative speed and piston speed for slow-speed 
engines are given in Fig. 39. 

Medium-speed engines generally operate at rotative 
speeds between 150 and 250 R.P.M. They are generally 
fitted with the better forms of multiported and balanced 
slide valves, with poppet valves, or with a positively operated 
Corliss type of valve. 

60. The Simple D-slide Valve Engine. The simplest 
and cheapest type of reciprocating steam engine manu- 
factured is shown in part section in Fig. 40 with the prin- 
cipal parts labelled. The cylinder, piston, steam chest 
and valve are sectioned in order to show the internal con- 
struction. 

This engine, like most steam engines, is double acting, 
that is, a cycle is produced on each side of the piston during 
every revolution. Steam is admitted and expanded on one 
side of the piston while steam is being exhausted on the other 
side. The control of admission and exhaust is effected by 
the slide valve and will be considered in detail in later sec- 
tions. 

The mechanical energy made available by the steam 
operating in an engine cylinder is not developed at a uniform 
or constant rate, but fluctuates, during each revolution, 
above and below the amount required to overcome the 
constant resistance at the shaft due to the work the engine 



THE REAL STEAM ENGINE 



97 



is doing. If no provision were made to prevent it, this 
would result in a very variable rate of rotation during 
each revolution. When the energy made available was in 
excess of the demand it would be used in accelerating 
the moving parts of the engine and the speed of the latter 
would increase. The reverse would occur when the supply 
did not equal the demand. 

The fly-wheel is used to prevent violent fluctuations 



Com'bined Flywheel 
and Beltwheel 




Fig. 40.— Simple D-slide Valve Engine. 

of this kind. It is made with a comparatively heavy rim 
and a great deal of energy must be supplied to accelerate 
it to any appreciable extent in a short time. Similarly 
it can give out a great deal of energy when slowing down. 
The fly-wheel therefore serves as a sort of reservoir in which 
excess energy can be stored temporarily and from which 
it can later be withdrawn when a deficiency exists. The 
fly-wheel thus acts as a damper to variation of rotative 



98 



STEAM POWER 



speed during each revolution, minimizing but not entirely 
eliminating such variation. It may also serve as a belt 
wheel, as shown in the illustration. 

The governor controls the steam supply to the cylinder 
in such a way that enough heat will be supplied to make 
available the power demanded at the shaft. Were more 
supplied the excess would be absorbed by the moving 
parts and the engine speed would increase, were less supplied 
the engine speed would decrease. 

61. Engine Nomenclature. The meanings of several 
terms used in describing engines are not self-evident, their 



Belt Backward 



Belt Forward 




Fig. 41. — Engine Nomenclature. 



definitions depending merely on accepted usage. Some 
of these terms and their meanings are illustrated in Fig. 41. 

The crank end of a horizontal engine is called the front 
of the engine, so that the cylinder head nearest the crank 
is called the front head and the stroke of the piston toward 
the crank is known as the forward stroke. The forward 
stroke of the piston is also spoken of as the outstroke, par- 
ticularly in connection with single acting engines. The 
stroke away from the crank is correspondingly designated 
as the return or the instroke. 

62. Principal Parts of Engines. The parts of engines 
may be roughly divided into stationary and moving, such as 
frame, cylinder, cylinder and valve chest covers, etc., which 



THE REAL STEAM ENGINE 



99 



are stationary, and piston, piston rod, crosshead, fly-wheel, 
etc., which are all moving parts when the engine is in opera- 
tion. The moving parts are often divided into reciprocal- 




Fig. 42. — Frame for Small Vertical Engine. 

ing and rotating parts. Thus the piston and all connected 
parts through and including the crosshead, and the valve 
and many connected parts in the case of slide valve engines, 



Jaws for 
main bearings 




Fig. 43. — Frame for Medium Speed Center Crank Engine. 



all reciprocate when the engine is in operation. The 
shaft, fly-wheel, eccentric sheaves and governor constitute 
the principal rotating parts. 



100 



STEAM POWER 



Some engines also have oscillating parts, such as the 
valves in Corliss types, which rock back and forth in the 
arc of a circle, and the rocker arms in various forms of 
valve gear, these arms rocking through a short arc about 
a fixed pin near one end. 

The principal parts and their functions are briefly con- 
sidered in the following paragraphs: 

(a) The Frame. The frame of the engine, sometimes 
known as the bed, serves to support the other parts, to tie 



.» Cylinder bolted 
1 to finished 
surface here. 




Jaws for 
main bearing 



Bored cross 
tieadguides 



Fig. 44. — Frame for Slow Speed Engine of Corliss Type; Side or 
Overhung Crank, 



them together in their proper relations and to fasten the 
whole structure to whatever foundation is used. The cross- 
head guides and the seats for the main bearings are incor- 
porated in the frame. 

The frame is commonly made of cast iron in the form 
of a hollow box which is properly ribbed to give the neces- 
sary stiffness. 

Examples of frames are shown in Figs. 42, 43 and 44. 



THE REAL STEAM ENGINE 



101 



(6) The Cylinder and Steam Chest. The cylinder and 
steam chest are generally incorporated in the same casting, 
and surfaces of covered cavities in this casting are finished 



SteamJPipe 




Steam dies 
Cover 



-Steam Chest 



Fig. 45. 

-Steam Chest Cover 



Steam 
Chest 




'Cylinder 
Fig. 46. 



to form the cylinder in which the piston operates and the 
seat or seats upon which valves rest and move. 

The cylinder may be single walled with flanges on the 
end to receive the cylinder head, as illustrated in Fig. 40 



102 



STEAM POWER 



(plain D-slide valve), in which case a thin sheet-metal 
jacket is fastened around it and the space between filled 
with heat-insulating material. Or, the cylinder may be 
cast with double walls, the space between the two being 
used as an air jacket or as a steam jacket. 

Some cylinders are fitted with a liner, which is a plain 
cylinder pressed into place within the cylinder casting 
and forming the bore of the working cylinder. This prac- 




Fig. 47. — Section of Atlas Medium Speed Engine, Showing Balanced 

Slide Valve. 

tice is common on the larger types, the liner being used so 
that when wear has occurred it can be replaced cheaply, 
instead of it being necessary to rebore or even replace 
the main casting. 

Examples of cylinder construction are shown in Figs. 
45, 46, 47, 48, 49 and 50. 

(c) The Piston. The function of the piston is two- 
fold. It must prevent the leakage of steam by it from one 
end of the cylinder to the other, and it must receive the 



THE REAL STEAM ENGINE 



103 




104 



STEAM POWER 



pressures exerted by the steam and transmit them to the 
other parts of the mechanism as it moves. 



Chambers for 

Steam 'Valves 

(admission valves) 



Chambers for 
Exhaust Valves 



Steam Port 




Fig. 49. — Corliss Cylinder Casting. 
Steam Connection 




Fig. 50. — Corliss Cylinder with Lagging in Place. 

Leakage of steam is prevented by the use of piston 
rings, which are metal rings fitted into grooves in the circular 
surface of the piston and pressed out against the cylinder 




THE REAL STEAM ENGINE 105 

Pistons 



Fig. 51. 





Piston Rings 



Piston Body 
or Spider 



^ 



Follower Plate 



4> 



O- 



J 



End of 
Piston Rod 



W////MM 

'/////// /fori ■ 

[Mm 




'Bull Rings 

Fig. 53. — Built-up Piston Used in Large Engines. 



106 



STEAM POWER 



walls by spring action. They may be made of one piece 
of metal turned into a ring of slightly larger diameter than 
the cylinder, cut through and sprung into place, or they may 
be made in pieces as shown in Fig. 51, and pressed out 
against the wall by small helical or leaf springs. 

The piston itself may consist of a solid disk of metal 
fitted with a hub and a short cylindrical part with grooves 
for the rings, as shown in Fig. 52 (a) and (b), or it may be 
an elaborate built-up structure, as shown in Fig. 53. 




Fig. 54. — Crosshead and Pin. 



(d) The Piston Rod. The piston rod is a plain circular 
steel rod fitted with such shoulders and threads at the ends 
as are necessary for the fastening of the piston and the cross- 
head. Examples of such fastenings are given in Figs. 
51, 52, 53 and 55. 

In large horizontal engines the piston rod sometimes 
extends through the piston and rear cylinder head, and 
the rear end is then supported by a small auxiliary cross- 
head. The extension of the rod is known as the tail rod 
and the auxiliary crosshead as the tail rod crosshead. 



THE REAL STEAM ENGINE 



107 



Such constructions are used when the weight of the piston 
is so great that it would cause serious cylinder wear if not 
supported more perfectly than is possible with the ordinary 
overhung arrangement. 



v ;; . .1 | 




1 l 






Fig, 55. — Single Slipper Crosshead. 




Fig. 56. — Crosshead with Adjustable Slippers. 

(e) The Crosshead and Guides. The crosshead and 
guides are used for the purpose of supporting the piston 
and its rod and guiding them in a straight line. The 
crosshead also serves to connect the piston rod and the 
connecting rod through which the forces are transmitted 
to the crank pin. 



108 



STEAM POWER 



Crossheads are generally cast in the form of imperfectly 
shaped boxes and carry slippers which are faced with anti- 
friction metal where they come in contact with the guides. 
The slippers may be flat and operate on planed guides, 
as shown in Fig. 40, or they may be turned and operate in 
bored guides, as shown in Figs. 48, 54, 55 and 56. Pro- 
vision is generally, though not always, made for taking 
up wear of guides and slippers by setting the slippers 
further out from the body of the casting. 

With the type shown in Figs. 40 and 54 this can be 






Fig. 57. — Solid End Connecting Rod; for Overhung Cranks only. 



done by the insertion of thin sheets of metal or paper 
(known as shims) between the body of the crosshead and the 
slippers. In the type shown in Fig. 56 the slippers are 
finished with inclined surfaces where they come in contact 
with the main casting, and the adjustment is made bj^ wedg- 
ing the slippers apart by the use of the adjusting bolts 
shown. 

The wrist-pin end of the connecting rod enters the 
crosshead casting and is held in place by means of the 
wrist pin, about which it oscillates when the engine is in 
operation. 



THE REAL STEAM ENGINE 



109 



(/) The Connecting Rod. This rod connects the re- 
ciprocating crosshead with the rotating crank shaft and 
transmits the forces from one to the other. It consists 
of a body or shank and two ends or heads. The ends may 




ujjiuj 



Fig. 58. — Connecting Rod with Bolted Strap Ends; May be Used 
with Center or Side Crank Constructions. 

be " closed " or " solid " as shown in Fig. 57; they may 
be made with a strap bolted in place as shown in Fig. 58; 
or the crank-pin end may be made of two half boxes bolted 
together to form a " marine end " as shown in Fig. 59. 
The ends are always made adjustable so that wear 



, Crank End 



Crosshead End. 




Fig. 59. — Marine End Connecting Rod. 



can be taken up, thus preventing noisy operation due to 
hammering between the ends and the pins at times when 
the direction in which forces act is reversed. With solid 
and strap types this adjustment is generally made by 
means of wedges similar to those shown in Figs. 57 and 58. 



110 



STEAM POWER 



With the marine type shims are used between the two 
halves, and the diameter of the hole formed by the latter 
is decreased by the removal of shims of the required thick- 
ness. 

(g) The Shaft. The crank shaft itself is generally made 




Fig 60.— Crank Shaft, Center Crank. 




Counterweights 
or Balances 

Fig. 61. — Center Crank. 




Counterweight 

Fig. 62.— Center Crank. 



of steel, but the counterbalances are often of cast iron. It 
may be one forging throughout or may be built up by 
shrinking the various parts together. Multicrank shafts 
of large size are generally of built-up construction, the crank 
pins being shrunk into the crank arms and the latter shrunk 
on to the pieces of shaft. 

The counterbalance weights are used to balance the 



THE REAL STEAM ENGINE 



111 



centrifugal effect of the crank pin, part of the crank arms 
and part of the connecting rod, all of which rotate off center. 
In some engines part of the unbalanced effect of the recip- 
rocating parts is also imperfectly balanced by these counter- 
balances. 

Various types of shafts are shown in Figs. 60, 61, 62, 
63, and 64. 

(h) Bearings. Bearings are distinguished as main and 
as outboard bearings. Main bearings are those carried by 




Disci"! 



Fig. 63.— Crank Shaft and Disc, Overhung Crank. 




Fig. 64. — Overhung Crank. 



the frame of the engine and outboard bearings are carried 
by separate pedestals or by pedestals fastened to a plate 
which is in turn fastened to the frame. Center-crank 
engines have two main bearings, and side-crank engines 
only have one, the other end of the shaft being supported 
by an outboard bearing. 

The bearings of steam engines are generally formed 
of babbitt-lined boxes carried within jaws machined in a 
frame, or in a separate pedestal, and held in place by a 
bearing cap. The boxes are made in two, three or four 
parts to allow for adjustment to compensate for wear and 



112 



STEAM POWER 



to give a certain degree of flexibility. Adjustment for 
wear is either made by means of wedges or by means of 
screws which force the various parts of the boxes toward 
the shaft. An example of a three-part bearing with screw 
adjustment as used with large side-crank engines is shown 
in Fig. 65. The parts of a three-part bearing with wedge 
adjustment are shown in Fig. 66. 

Bearings are often lubricated by rings or chains, and 
they are then known as ring- or as chain-oiling bearings. 




Fig. 65. — •Three Part Main Bearing with Screw Adjustment. 



In the ring-oiling bearing one or more metal rings of large 
diameter hang loosely on the shaft within the bearing and 
dip into an oil reservoir below the shaft. Rotation of the 
shaft causes the rings hanging on it to rotate and they 
carry oil up from the reservoir and spill it out over the shaft 
within the bearing. Chain bearings are similar except that 
chains are substituted for rings. 

(i) Fly-wheels. The function of the fly-wheel has al- 
ready been considered and need not be discussed further. 
The wheel is constructed with a heavy rim joined to a hub 



THE REAL STEAM ENGINE 



113 



by six or eight arms. In the smaller sizes the wheel may be 
cast in one piece, but best practice calls for a split hub in 
that case to partly equalize certain casting strains which 
result from unequal thicknesses of metal in different parts 
of the wheel. Large wheels are cast in two or more parts 
both for the purpose of partly avoiding casting strains 
and for the purpose of facilitating handling and shipping. 





Fig. 66. — Three Part Bearing 
Showing Wedge Adjustment. 



Fig. 67. 



A two-part wheel with the rim sections joined by 
prisoner links shrunk in place and the hub fastened with 
bolts is shown in Fig. 67. 



PROBLEMS 

1., A given engine has a piston displacement of 3 cu.ft. and a 
clearance volume of 3%. Compression begins when 85% of the 
exhaust stroke has been completed and the pressure within the 
cylinder at that time is 16 lbs. per square inch absolute. Deter- 
mine the weight of the cushion steam on the assumption that this 
steam is dry and saturated at the beginning of compression. 



114 STEAM POWER 

2. Assume the engine described in Prob. 1 to cut-off at \ stroke 
and with a pressure inside the cylinder equal to 115 lbs. per square 
inch absolute. Find the weight of cylinder feed if the quality 
of the material in the cylinder at the time of cut-off is 75%. 

3. Find the piston speed of an engine with a stroke of 2 ft. 
and a rotative speed of 150 R.P.M. 

4. Show by means of Heck's formula that initial condensation 
increases with pressure range. 



CHAPTER VIII 
THE INDICATOR DIAGRAM AND DERIVED VALUES 

63. The Indicator. The ideal steam engine cycle was 
described in Chapter IV, and the sort of diagram which 
would be obtained from a real engine was shown in Chapter 



Drum 



Point Holder 



Cylinde 




Fig. 68. 

VII; but the means by which such diagrams are obtained 
from operating engines was not given. 

Indicator diagrams showing the pressure and volume 
changes experienced by steam in the cylinders of real 

115 



116 



STEAM POWER 



engines are obtained by means of an instrument known as 
an indicator. The operation of obtaining such diagrams is 
known as indicating the engine. 

An external view of one form of indicator is shown in 
Fig. 68 and a section through the instrument is given in 
Fig. 69. The method of connecting an indicator to the 



Piston Rod 




"Connected to Engine 
Cylinder 



Fig. 69. 



cylinder of a steam engine and one method used for driving 
it are illustrated in Fig. 70. 

The indicator is intended to draw a diagram showing 
corresponding pressures and volumes within the engine 
cylinder and must, therefore, contain one part which will 
move in proportion to pressure variations and another which 
will move in proportion to volume changes. The one may 



INDICATOR DIAGRAM AND DERIVED VALUES 117 

be called the pressure-measuring and the other the volume- 
measuring device. 

The pressure-measuring device generally consists of 
a piston, such as shown in the figure, working with minimum 
friction in a small cylinder and fitted with a spring which 
will resist what may be called outward motion (upward 
in the figure). The cylinder containing this piston is 
coupled to a short pipe connected with the clearance space 
of the engine and, whenever the indicator cock in this 
connection is open, the steam acting on the engine piston 




Fig. 70. — Method of Attaching and Operating an Indicator. 



will also act on the indicator piston. Steam of any given 
pressure will drive the indicator piston out against the 
action of the spring until the pressure exerted by the spring 
is equal to that exerted on the face of the piston. The 
indicator piston will thus move out different distances for 
different pressures, and, through the piston rod and pencil 
mechanism, will move the pencil point to various heights 
corresponding to different steam pressures. The pencil 
mechanism is so arranged that the point traces a straight 
vertical line on the drum as the indicator piston moves in 
and out. 

Springs are made to certain definite scales, thus there 



118 STEAM POWER 

are, for instance, 10-lb., 25-lb., 50-lb. and 100-lb. springs. 
The number which is known as the scale of the spring 
designates the steam pressure in pounds per square inch 
which is required to move the pencil point 1 inch against 
the action of such a spring. With a 100-lb. spring in the 
indicator, a steam pressure of 50 pounds per square inch 
acting on the indicator piston would drive, the pencil up a 
distance of half an inch, a pressure of 100 pounds per square 
inch would give 1 inch of motion and so forth. 

The volume-measuring device is of an inferential kind. 
It simply indicates the position attained by the engine 
piston at the time when a given steam pressure existed in 
the cylinder and the volume occupied by the steam can be 
calculated from piston position and cylinder dimensions. 
The position of the piston is indicated by connecting the 
cord wound around the drum to some part of the engine 
which is rigidly connected to the piston. The crosshead 
is commonly used for this purpose and, since the motion 
of this member is generally much greater than the circum- 
ference of the drum, it is necessary to use a reducing 
mechanism of some sort. This mechanism must be very 
accurate, so that it moves the drum as nearly as possible 
in proportion to the motion of the engine piston. 

The pencil point moves up and down as the pressure 
within the cylinder varies, and the drum rotates under 
the point in proportion to the motion of the engine pis- 
ton, so that the combination of the two motions brings the 
pencil point to successive positions on the drum which indi- 
cate successive corresponding values of steam pressure and 
piston position. By mounting a piece of paper, known 
as a card, on the drum and pressing the pencil point upon 
this paper, the successive positions occupied by the pencil 
point will be recorded in the form of a series of curves and 
straight lines. 

If the drum is rotated with the lower side of the indica- 
tor piston connected to atmosphere, the pencil will trace 



INDICATOR DIAGRAM AND DERIVED VALUES 119 

a horizontal line. This is known as the atmospheric line 
and is used as a reference for locating the pressure scale. 
If the indicator cylinder is then connected with the engine 
cylinder and the drum is rotated by the reducing mechanism, 
a diagram similar to that of Fig. 71 will be drawn upon the 
card. The atmospheric line indicates the height assumed 
by the pencil when atmospheric pressure acts on the 
piston and, knowing the value of the existing atmospheric 
pressure (barometer reading) and the scale of the spring, 
a line at a height representing zero pressure can be drawn 
on the card. This line is indicated in Fig. 72. 

The length of the card between the lines a and b is 
proportional to the length of the engine stroke and there- 




Atmospheric Line 



Fig. 71. 







a 




h) 






















Sj _ 


> 






1 


w - 


a 






1 




N 








a ■ 


a 


















'/. 










o 


J 








Ph - 






^ f Atmospheric Line 














Line of Zero Pressure \ 



Volume Scale 



Fig. 72. 



fore to the piston displacement, that is, to the volume 
swept through by the piston. Knowing the clearance 
volume of the engine as a percentage or fraction of the 
piston displacement, this fraction of the length of the 
diagram can be laid off from the end of the diagram to 
give a line of zero volume. This line is also indicated in 
Fig. 72. 

With the line of zero pressure and the line of zero 
volume drawn in, all values of steam pressure and volume 
occupied by steam can be read directly from the diagram, 
and it thus forms a picture of what occurs within the real 
engine cylinder. 

The indicator diagram is used for a number of purposes, 
the more important being: 



120 



STEAM POWER 



Length of Diagram — >t 



(1) The determination of the energy made available 
within the cylinder, that is, the indicated horse-power, I.h.p. 

(2) The determination of the amount of initial conden- 
sation and of heat interchanges between walls and cylinder. 

(3) The determination of what is known as the diagram 
water rate. 

(4) The study of the operation and timing of valves. 
The second one of these uses has already been considered 
in Chapter VII, the others are treated in succeeding sections. 

64. Determination of I.h.p. The lines of the indicator 
diagram show by their height the pressures or forces acting 
on the engine piston as it moves. But the product of force 
by distance is equal to work and these lines can be used there- 
fore for determining the net 
work done by the steam 
upon the piston. 

In Fig. 73 is shown the 
upper part of the diagram, 
the curved lines represent- 
ing the successive pressures 
in pounds per square inch 
which acted on the left face 
of the piston while it moved 
outward. If the average 
pressure could be deter- 
mined and multiplied by 
the area of the piston face, this product would be the 
average total force acting on the piston. Multiplying this 
by the distance traveled would give the work done by 
the steam upon the piston. Expressed in the form of an 
equation, 

#o = PoXaXLft.-lbs., (33) 

in which 

Z?o = work done upon piston by steam during outstroke; 
po = mean pressure (in pounds square inch) acting on 
piston during outstroke; 









i i 


*■ i i 




5 




V i i 


i i 
i i 



:-:-^ 



Fig. 73.— Positive Work Area. 






INDICATOR DIAGRAM AND DERIVED VALUES 121 

a = area of piston face in square inches; and 
L = stroke of piston in feet. 

For the instroke shown in Fig. 74, the work done by 
the piston on the steam is given by the similar expression _ 



E i = p i XaXLitAbs., 



. (34) 



= A, 



$$$$^ v 



I 



-=3- 



Fig. 74. — Negative Work Area. 

in which E t and p% represent work done and mean pressure 
respectively. 

The net work done by the steam upon the piston per 
cycle is then, 



#cycie = Eo — Ei = (p — pi)aL ft. -lbs. 



(35) 



The values of p and p t can be found directly from the 
diagram by dividing the areas A and A t respectively by the 
length I and then multiplying by the scale of the spring, 
giving 

Po = -7-Xscale of spring, 
and 

Pi = -y i X scale of spring, 



122 STEAM POWER 

so that, 

Vo~Vi~V = ~^i — *X scale of spring . . . (36) 

= areaof f agram Xscale of spring. . (37) 

The value of p evidently can be determined very simply 
from the indicator diagram, and the work per cycle can be 
found when p is known by substituting in the following 
equation, obtained by putting p for po — pi in Eq. (35), 

#cycie = pXaXL ft.-lbs (38) 

The pressure p is known as the mean effective pressure and 
is often represented by M.E.P. 

If n cycles are produced per minute, the net work done 
by the steam upon the piston per minute will be 

E m m = pXaXLXn, (39) 

which is generally rearranged to read, 

E m m = pLan, (40) 

in which form the group of letters forming the right-hand 
member is easily remembered. 

Since 33,000 foot-pounds per minute are equivalent 
to one horse-power, it follows that the power made avail- 
able as shown by the indicator diagram, that is, the indicated 
horse-power, must be, 

,, pLan 

Lhp - =337)00 (41) 

in which 

p = mean effective pressure in pounds per square inch; 

L = stroke of piston in feet ; 

a = area of piston in square inches; 

= (diam. cyl. in inches) 2 X7r/4 = .7854d 2 ; and 
n = number of cycles per minute. 






INDICATOR DIAGRAM AND DERIVED VALUES 123 

If an engine cylinder takes steam on one side of the 
piston only, that is, if the cylinder is single acting, the num- 
ber of cycles produced per minute is equal to the number 
of revolutions per minute, but it should be noted that for 
other arrangements this is not necessarily true. In the 
case of double-acting engines which receive steam at both 
ends of the cylinder, the number of cycles produced is equal 
to twice the number of revolutions. 

It should also be noted that the symbol a represents 
the area of the piston face upon which the steam acts. 
If a piston rod extend from the face of the piston to and 
through the cylinder head (as is always the case at the 
crank end of double-acting cylinders), the area of the 
piston rod must be subtracted from that of the piston to 
obtain the area on which the steam really acts. When a 
tail rod is used, a correction must be made for each side 
of the piston. 

In the case of double-acting engines the indicated 
horse-power may be determined in two ways: It may be 
figured separately for the two ends of the cylinder, or 
the values for the area and pressure may be averaged for 
the two ends and the value of n chosen equal to twice 
the revolutions per minute. The former is generally the 
more accurate method. 

It will have been observed that the area of the indi- 
cator diagram must be determined before the mean 
effective pressure can be found. This area is generally 
measured by means of an instrument known as a planimeter, 
and this is the most accurate method. It occasionally 
happens, however, that a planimeter is not available when 
the value of the indicated horse-power is desired. Under 
such circumstances an approximate determination of the 
area of the indicator diagram can be made by the method 
of ordinates. 

For this purpose the length of the diagram is divided 
into an equal number of parts, usually ten, as shown in 



124 



STEAM POWER 



Fig. 75 and vertical lines 2/1, 2/2, 2/3 > etc., are drawn at the center 
of each of the parts into which the diagram has been divided. 
The mean ordinate or height is then found from the equa- 
tion, 

V\ +2/2 +2/3 +1/4+ etc. 
number of vertical lines' 



(42) 



and the mean effective pressure is then determined by- 
multiplying 2/m by the scale of the spring. 

An indicator diagram similar to that shown in Fig. 76 
is occasionally obtained. The small loop on the end repre- 
sents negative work, since the pressure of the steam which 





Fig. 75. 



Fig. 76. 



does work upon the piston is lower than that which resists 
the return of the piston. When using a planimeter, this area 
is automatically subtracted from that of the rest of the 
diagram, but care should be taken to see that this is also 
done when the method of ordinates is used. 



ILLUSTRATIVE PROBLEM 

1. Determine the I.h.p. of a double-acting steam engine, having 
a cylinder 8 ins. diameter, length of stroke, 12 ins., running at 
100 R.P.M., the mean effective pressure (M.E.P.) on the piston 
being 45 lbs. Neglect the area of the piston rod. 



I.h.p. = 



pLan (pXa) lbs.X(Ln) ft. per min. 



33,000 33,000 ft.-lbs. per min. 

(45X8X8X.7854) lbs. XjfX 100X2) ft. per min. 
33,000 



INDICATOR DIAGRAM AND DERIVED VALUES 125 

2260 lbs. X 200 ft. per min. 
33,000 

= 14 nearly. 

2. The I.h.p. of a double-acting engine is 14, the R.P.M. =100; 
M.E.P. =45 lbs.; length of stroke = 12 ins. Find the diameter 
of the cylinder, neglecting area of piston rod. 

First determine the area of the piston from the formula 

pLan 33,000 I.h.p. 

Lh - P - =3^000 ° r Q= pXLXn 5 

33,000X14 ri , . wd* 

a= 45^naoo^2 =5l ' 4sq - m - = T ; 

/ £1 A , 

d = A / ' = V 65.4 = 8 ins. (approx.) . 
\ . i bo4 

65. Conventional Diagram and Card Factors. It is 

often necessary to approximate the mean effective pressure 
obtained in the cylinder of an engine when no indicator 
diagrams are available. The most common case is when 
an engine is being designed to carry a certain load and it 
is desired to determine the necessary cylinder dimensions 
and speed. If the probable mean effective pressure can be 
determined, the dimensions and speed can be found from the 
equation, 

pLan 



I.h.p. per cylinder end 

oo, (JUL) 



by rewriting it 
from which 



_ pLn X0.7854r/ 2 
l.p.n.- 33()00 



Since n is equal to revolutions per minute for one cylinder 
end, the product of L by n must be equal to half the piston 



126 STEAM POWER 

speed of the engine, and a proper value of this product can 
be chosen for substitution in the equation. If a proper 
value for p can then be predicted the only unknown remain- 
ing will be the diameter d, and this can be found by solving 
the equation. 

The prediction of the mean effective pressure is made 
either by drawing upon recorded experience in the form 
of values obtained in similar engines previously constructed 
or by means of what is known as a conventional indicator 
diagram. 

The conventional diagram is drawn with upper and 
lower pressures equal to those expected in the case of the 
real engine, and all expansions and compressions are drawn 
as rectangular hyperbolas. The equation of the rectangu- 
lar hyperbola is 

PlVl =P n V n , (44) 

in which subscript 1 indicates initial conditions and sub- 
script n represents any later conditions with the same mate- 
rial in the cylinder. This law is assumed because it is the 
simplest and, as a rough average, gives values as close to 
those actually attained as do any of the more complicated 
laws. 

The diagram may be drawn as nearly as possible like 
the one which the engine may be expected to give or it 
may be drawn with various simplifications which remove 
it more and more from the approximation to an actual 
indicator diagram. In any case, the mean effective pres- 
sure is determined from this diagram and this value is then 
multiplied by a corrective factor, the value of which has 
been determined by experience. This corrective factor is 
called the diagram factor or card factor and it is really 
the ratio of the area of the diagram the engine would 
really give to the area of the conventional diagram 
used. 

The simplest form of conventional diagram is drawn 



INDICATOR DIAGRAM AND DERIVED VALUES 127 



by neglecting the clearance volume and has the shape shown 

in Fig. 77. The upper line is drawn horizontal at a height 

representing the highest pressure expected and of such a 

length (compared with the length of the diagram) as will 

approximately represent the fraction of the stroke at which 

cut-off is to occur in the real 

engine. The expansion curve is 

then drawn in as a rectangular 

hyperbola and extended until the 

end of the diagram is reached. 

The next line is drawn vertical 

and the lower line of the diagram 

is drawn horizontal at a height 

representing the pressure expected 

in the space into which the engine 

is to exhaust. 

This simple diagram can be 
divided into the three areas shown 

and the value of the work represented by these areas can be 
determined from the equations given below, the first and 
last of which should be self-evident from what has preceded. 
The equation for the work represented by area A 2 can be 
determined very easily by means of integral calculus. The 
equations are, 




Fig. 77. — Conventional In- 
dicator Diagram. 



and 



Ai represents P1V1 ft. -lbs.; 



A2 represents P\V\ \og e -rf- = PiV\ log e r ft.-lbs., 
v 1 



A3 represents P2V2 ft.-lbs. 



in which P represents pressure in pounds per square foot and 
V represents volume in cubic feet. 

The total area is then equal to the sum A1+A2 — A3 
and the net work is equal to a similar sum of the right- 



128 STEAM POWER 

hand members given above. The net work must also 
equal the mean-effective pressure P m multiplied by the 
total volume change, so that 



and 



PmV 2 = PiVi+PiV 1 \oger-P 2 V2, . . (45) 



Pm = Pi^r+P^ loge r-P 2 . . . (46) 

V 2 V 2 

= p '(f:+ft io ^)- p2 ' • • • <«> 

1 V 

and substituting — for ■=- this gives 
& r V 2 

P.=Pi(i±^)-P a (48) 

V 2 
The ratio -ry~ = r is called the ratio of expansion and its 

v 1 

Vi 1 . 
reciprocal, r— = — is known as the cut-off ratio. By the use 

of this ratio the volume terms can be disposed of and the 
equation above is obtained. This equation then gives the 
mean effective pressure in terms of upper and lower pres- 
sures and the fraction of the stroke at which cut-off is 
desired in the real engine and no cylinder dimensions need 
be known. 

Since pressures in steam-engine practice are usually 
given in pounds per square inch, the equation for mean 
effective pressure is more useful in the form 

Pm = pi (\±l2!kL\ p2> .... (49) 



in which pi and p 2 and p m are expressed in pounds per 
square inch absolute. For convenience in the use of this 



INDICATOR DIAGRAM AND DERIVED VALUES 129 

equation the values assumed by the bracketed quantity 
are given for various conditions in Table III. 



TABLE III 



r 


1 +log e r 
r 


r 


1 +log e r 


r 


1 +log e r 




r 


r 


1.0 


1.00 


6.0 


0.465 


16.0 


0.236 


1.5 


0.937 


7.0 


0.421 


17.0 


0.226 


2.0 


0.847 


8.0 


0.385 


18.0 


0.216 


2.5 


0.766 


9.0 


. 355 


19.0 


0.208 


3.0 


0.700 


10.0 


0.330 


20.0 


0.200 


3.5 


0.644 


11.0 


0.309 


21.0 


0.192 


4.0 


0.597 


12.0 


0.290 


22.0 


0.186 


4.5 


0.556 


13.0 


0.274 


23.0 


0.180 


5.0 


0.522 


14.0 


0.260 


24.0 


0.174 


5.5 


0.492 


15.0 


0.247 


25.0 


0.169 



The values of the mean effective pressures obtained 
from this form of diagram are very much higher than are 
to be expected from real engines with the same initial and 
terminal pressures and the same nominal ratio of expan- 
sion. They are therefore corrected by multiplying by the 
proper diagram factor as selected from Table IV. It is 
obvious from the range of values given that the selection 
of a proper value for the factor depends largely on expe- 
rience, but such experience is quickly gained by contact 
with real engines and a study of the practical diagrams. 

TABLE IV 
Diagram Factors 

Simple slide-valve engine 55 to 90% 

Simple Corliss engine 85 to 90 

Compound slide-valve engine 55 to 80 

Compound Corliss engine 75 to 85 

Triple-expansion engine 55 to 70 

66. Ratio of Expansion. — The ratio of expansion used 
above is sometimes called the apparent ratio. It is not the 



130 



STEAM POWER 



real ratio of expansion for an engine with clearance, 
such an engine the real ratio of expansion is 



For 



V2 + Va 



(50) 




Fig. 78. 



in which the symbols represent 
the volumes indicated in Fig. 
78. 

The numerical values of r 
and r' are often very different 
and care should be used in dis- 
tinguishing between them. The 
diagram factors referred to in 
Table IV are for idealized con- 



ventional cards without clearance as shown in Fig. 77. 



ILLUSTRATIVE PROBLEMS 

1. Given an engine with a stroke of 24 ins. and cut-off occurring 
at \ stroke. Steam pressure of 160 lbs. per square inch and 
back pressure of 16 lbs. Assume diagram factor =80%. Neglect- 
ing clearance, find the probable M.E.P. 



M 



KP ' =P \ r~~)~ P2 = 



160 



l+log e 3 



-16 



= 160X.7-16 = 112.0-16=96 1bs. 



Hence probable M.E.P. =.80X96 =76.8 lbs. 

2. A given double-acting engine indicates 75 I.h.p. under the 
following conditions: 

Cut-off at 20%; steam pressure, 140 lbs. per square inch 
absolute; piston speed, 600 ft. per minute; back pressure, 2 lbs. 
per square inch absolute. 

Assume a diagram factor for this type of engine equal to 85%; 
and neglecting clearance, find a convenient size of the cylinder 
(diameter and stroke). 



INDICATOR DIAGRAM AND DERIVED VALUES 131 
Solution. 

r= Jo =5; 

Pm=p (^ftlj _ p , =140 (kfclli^ _ 2 =140( 522) _ 2 

= 73.1 -2=71.1 lbs. per sq.in. 

Diagram factor =85%. Hence probable 
M.E.P. =71.1 X.85 =60.4 lbs. 

Therefore, since 

T1 pLan 75X33,000 ^ . 

Lh - P ' = 3P00 ° = 604^60^ = 68 - 3sq ' m ' 

d = 9^ ins. (approx.) ; 
and since 2Ln = 600, assume L = 1 ft. 

hence ?i=300R.P.M. 

The engine is rated 9.5X12 ins., running at 300 R.P.M. 

67. Determination of Clearance Volume from Diagram. 

It was shown in a preceding paragraph that the clearance 
volume of a cylinder must be known in order to draw the 
line of zero volumes on the indicator diagram. This 
volume can be determined accurately for any real engine 
by weighing the quantity of water required to fill the clear- 
ance space, but this procedure is often impossible and 
an alternative, though approximate, method is often 
resorted to. 

This method is graphical and depends upon the assump- 
tion of the law of expansion and compression. As in the 
case of the conventional diagrams, expansion and compres- 
sion are assumed to follow rectangular hyperbolas. 

It is a property of this curve that diagonals such as 
aa and bb drawn for rectangles with their corners on the 



132 



STEAM POWER 



curve all pass through the origin of coordinates as shown 
in Fig. 79. 

If two points a and c are selected on the expansion 
curve of a real diagram and a rectangle is drawn upon 
them as shown in Fig. 80, the diagonal bd extended will 
pass through the origin of coordinates, if the expansion 
follows the assumed law. The point at which this diagonal 
cuts the zero pressure line must therefore be the point 
through which the vertical line of zero volume is to be drawn. 

If the original assumption were correct, this construc- 
tion would give the same point when different locations 
of the points a and c were chosen and when used on the 




Fig. 79.— Rectangular Hyperbola. 



1 

1 


-W 


> 

a 


-%r 


^Wl 






^ 


c 




I 






& 



Fig. 80. 



Atm^ 



compression as well as on the expansion line. In reality 
it will generally give as many different locations for the 
origin as are chosen for the rectangle abed. It is customary 
to construct this rectangle of fair size and to locate it near 
the center of the expansion curve. 

68. Diagram Water Rate. As was shown in an earlier 
chapter, part of the steam supplied an engine is generally 
condensed upon the cold metal walls surrounding it. The 
indicator diagram therefore shows the volumes assumed by 
the mixture of steam and liquid water in the cylinder, 
but, since the volume occupied by the liquid is negligible, 
it may be assumed to show the volumes occupied by the 
part of the mixture which exists in vaporous form. 



INDICATOR DIAGRAM AND DERIVED VALUES 133 



Assuming that the vapor is saturated, the volume 
occupied by one pound at various pressures can be found 
from the steam tables and, therefore, the weight existing 
in the cylinder can be calculated. The weight of steam 
determined in this way is known as the indicated steam/ 
the diagram steam or the diagram water rate. 

The diagram water rate is generally determined for a 
point such as z in Fig. 81 just 
after cut-off, though some 
engineers prefer to use a point 
nearer the lower end of the 
expansion curve. The volume 
occupied by the steam con- 
tained in the cylinder at point 
z is equal to V z and its weight 
can be determined by dividing 
this volume by the specific volume V 2 for the existing 
pressure P z . Thus, calling the weight of steam in the cyl- 
inder w z , 

y 

W * = W ( 51 ) 




Fig. 81. 



This quantity of steam is a mixture of cylinder feed 
and clearance or cushion steam and the weight of the latter 
must therefore be subtracted from w z to obtain the weight 
of cylinder feed w f . Assuming the cushion steam dry and 
saturated at the point k, the weight of cushion steam is 



Wb = 



W 



(52) 



so that the weight of cylinder feed per cycle as shown by 
the diagram at the point z is 

V z V k 



w f =w z -wi = = r - 

V 2 Vt 



(53) 



The formula is generally modified to give the steam 
consumption per indicated horse-power hour, instead of 



134 STEAM POWER 

per cycle, and it is also expressed in different terms as a 
matter of convenience. 
For this purpose let 

• y c i = clearance volume divided by piston displacement per 
stroke 

-Is. 

h' 

2/2 = piston displacement to point z divided by piston dis- 
placement per stroke 

Jz 

V 

y k = piston displacement to point h divided by piston dis- 
placement per stroke 
Jt 

V 

a = area of piston in square inches ; 

p = mean effective pressure in pounds per square inch; 

L = stroke in feet ; and 

n = number of cycles per minute. 

The piston displacement is then — -L cubic feet and 

the volumes at z and k are given by 

F.-(»Xxs)+(wXl5), 



and 



v *-(» x m)+(y<xm)' 



Substituting these values in Eq. (53) gives the cylinder 
feed per cycle as 

_ aL ( y+yg y k +y c l \ , 

Multiplying by the number of cycles per hour (60 Xw) 
and dividing by the indicated horse-power, £ gives 

oo,UUU 



INDICATOR DIAGRAM AND DERIVED VALUES 135 

the diagram water rate, or steam shown by the diagram per 
I.h.p. hour as 

_ 13,750 ( j/z+ya yt+vA rw 

w a — — ^ ^— J, . . . (o5) 

in which form the equation involves only values which can 
be determined directly from the diagram without any 
knowledge of the engine dimensions. 

The value obtained for w d will vary as the location of 
points z and k are varied because of the quality changes 
occurring during expansion and compression, and it is 
obvious that the diagram water rate is in no sense an accu- 
rate measure of the real water rate of the engine. It is, 
however, very useful for comparison with the real water 
rate, the ratio giving an indication of the loss by con- 
densation. 

Average values for real water rates are given in Chapter 
XL 



ILLUSTRATIVE PROBLEM 

Given the diagram shown in Fig. 82 and the following data 
from an actual test, find the diagram water rate for point c, and 
for point n. Double-acting steam 
engine having: 

Average piston area =28.9 sq.- 
in.; 

Length of stroke =8 in.; 

R.P.M. =237; I.h.p. =8.75; 
M.E.P. =31.6 lbs.; 

Clearance =13% ; Beginning of 
compression =29%; 

Weight of condensate per hour 
=371 lbs.; 

Quality at throttle =95%; 

Sp. vol. at c=7.8; 

Sp. vol. at n =12.57; 

Sp. vol. at K =38.4 cu.ft. per lb.; 

Assume .T£=100%. 




Fig. 82. 



136 STEAM POWER 

Solution. Substitution in Eq. (55) gives 
13,750 (yc+ya yk+ya\ 

13,750 /0.39+0.13 0.29+0.13 N 



31.6 \ 7.8 38.4 / 

= 24.2 lbs. per I.h.p. per hour at point c. 
13,750 fyn+yci yt+ya 



13,750 / 0.638+0.13 _ 0.29+0.13 \ 
" 31.6 V 12.57 38.4 / 

= 21.83. lbs. per I.h.p. per hour at point n. 



Real water rate = -^-X0.95 =40.2 lbs. 
8.75 

69. T^-diagram for a Real Engine. In Chapter VI 
the T^-diagrams of the various ideal cycles were given and 
attention was called to the fact that these diagrams were 
particularly useful, because they showed certain things 
which were not apparent from the more common PV- 
diagrams. 

It has been customary for many years to draw T<f>- 
diagrams for real engines by " transferring " the PV- 
diagram to T<£-coordinates, and various analytical and 
graphical methods have been developed for this purpose. 
There are certain unavoidable errors in all the methods 
used for drawing these diagrams, and the expansion curve 
is the only one of all the lines finally obtained which has 
any claim to accuracy. Even this curve is generally incor- 
rectly interpreted, because a knowledge of the exact weight 
of clearance steam is necessary for an accurate interpreta- 
tion and such knowledge is never available. 

Under the circumstances it seems unnecessary to con- 
sider in this book the rather complicated details involved 
in the construction of T0-diagrams purporting to show 
the behavior of steam in real engines. 



INDICATOR DIAGRAM AND DERIVED VALUES 137 

70. Mechanical and Thermal Efficiencies. The method 
of obtaining the indicated horse-power from the indicator 
diagram has been given in preceding paragraphs. In the 
real engine this power is not all made available at the shaft, 
because some of it is used in driving the engine against 
its own frictional resistance. Calling the power lost in 
this way the friction horse-power, it follows that 

Lh.p. = F.h.p.+D.h.p, .... (56) 

in which 

I.h.p. = indicated horse-power determined from the real 

indicator diagram; 
F.h.p. = friction horse-power, i.e., power required to 

drive engine; and 
D.h.p. = developed horse-power, i.e., power made avail- 
able at shaft. 

The developed horse-power is therefore always less 
than the indicated horse-power. The better the construc- 
tion of the engine the smaller the friction loss, and the 
measure of this loss is usually given in the form of an ef- 
ficiency. It is called the mechanical efficiency, and is defined 
by the equation 

Mech. eff. = ?M- (57) 

I.h.p. 

Values of mechanical efficiency range from about 80 per 
cent in the case of poorly designed and poorly adjusted 
horizontal engines to about 95 per cent in the case of the 
best vertical designs. 

The efficiency determined by dividing energy made 
available by heat supplied is known as the thermal efficiency. 
There are two possible thermal efficiencies, one based on 
the indicated power and the other on the developed power. 
The former is called the thermal efficiency on the indicated 
horse-power or the indicated thermal efficiency; the other 



138 STEAM POWER 

is known as the thermal efficiency on the developed horse- 
power or the developed thermal efficiency. Obviously 

Dev. ther. eff. = Mech. eff.X Indie, ther. eff. . . (58) 

The heat supplied may be assumed in two different 
ways; it may be taken as the total heat above 32° F. in the 
steam supplied the engine, or it may be taken as this value 
less the heat of the liquid corresponding to exhaust tem- 
perature. The second method is preferable, since it is 
reasonable to assume that the exhaust steam can be con- 
densed to water at the same temperature and that this water 
can be pumped to the boiler with the heat of the liquid 
corresponding to this temperature. This is practically 
parallel to the assumption made in treating the theoretical 
cycles. 

The thermal efficiencies are then 



Indie, ther. eff. 

I.h.p.X2545 



Heat above exhaust temp, supplied per hour' 

and 

Dev. ther. eff. 

= D.h.p.X2545 

Heat above exhaust temp, supplied per hour* 



(59) 



(60) 



Values of the indicated thermal efficiency range from 
about 5 per cent in ordinary practice with small engines to 
about 25 per cent in the best large engines. Values as low 
as 1 per cent are not uncommon with small, poorly designed 
and poorly operated engines. 

The actual performance of the cylinder of an engine 
is sometimes compared with the ideal possibilities as indi- 
cated by the Clausius and the Rankine cycles. The ratio 
of the work obtained in the real engine to that which could 
be obtained from the same quantity of heat with a Rankine 



INDICATOR DIAGRAM AND DERIVED VALUES 139 

or Clausius cycle is a measure of the performance of the 
real cylinder. This ratio is variously designated as cylinder 
efficiency, indicated efficiency, relative efficiency, etc. Its 
values range from less than 40 per cent to over 80 per cent, 
the highest recorded value being just over 88 per cent. 

PROBLEMS 

1. Using Table I, Chapter I, plot the specific heat of water 
between the range of temperatures of 20° F. and 300° F. for the 
intermediate values given. By the ordinate method for finding 
the mean height of an indicator diagram, determine the mean 
or average specific heat over this range. 

2. A double-acting engine Is required to give 50 I.h.p. under 
the following conditions: 

Cut-off =25%; 

Steam pressure =150 lbs. per square inch absolute; 
Back pressure = 16 lbs. per square inch absolute; 
Piston speed =540 ft. per minute. 

If the diagram factor for this type of engine is 75%, find the 
diameter of the cylinder and select the stroke and R.P.M. 

3. Assume a single-acting engine with 10-in. diameter and 12-in. 
stroke, 10X12 ins., to have cut-off occur at various points between 
10% and 50% of stroke. Assume also the pressures, speed, and 
card factor as given in Prob. 2. Find the probable I.h.p. at 
different cut-offs. 

4. Given an 18X.24-in. engine running at 120 R.P.M. 
Back pressure =2 lbs. per square inch absolute; 
Clearance = 10%; 

Cut-off =40%; 

Diagram factor =85%. 

Supposing* cut-off to remain constant, find the I.h.p/s cor- 
responding to steam pressure of 50, 90, and 130 lbs. per square 
inch absolute. 

5. Find the weight of dry steam which must be supplied per 
I.h.p. hour for each case of the previous problem, assuming the 
quality at cut-off to be 80%. Assume compression pressure to 
be 30 lbs. absolute and that steam is dry and saturated at end of 
compression. 

6. Find the quality of steam at cut-off in a cylinder, in which 
the piston displacement is 0.1278 cu.ft.; clearance = 10%; cut-off 
at 25% stroke; steam pressure at cut-off, 115 lbs. per square 



140 STEAM POWER 

inch absolute, and weight of steam in the cylinder at cut-off = 
0.012 lb. 

XT , A ri Actual vol. 

JNote. Quality == . . , — - for the given pressure. 

Weight XSp. vol. 

7. The piston displacement of a certain engine is 0.2 cu.ft. 
What weight of steam is in the cylinder at release where quality 
is 90%, and pressure is 25 lbs. per square inch absolute, if the 
clearance is 10%, and release occurs at 95% of the stroke? 

8. Find the weight of cushion steam in a 6X6 in. engine in 
which clearance = 15%; compression begins at 85% of the return 
stroke; back pressure is 14.7 lbs. per square inch absolute, and 
the quality of the cushion steam at the beginning of compression 
is 95%. 

9. Find the pressure and quality at the end of the compression 
line of the previous problem, assuming it to be adiabatic. 

10. An 8X10 in. engine running at 300 R.P.M. is double-acting, 
and cuts off at 15% of the stroke at a pressure of 120 lbs. per 
square inch absolute. It has a steam consumption of 35 lbs. per 
I.h.p.-hour. The compression begins at 60% of the return stroke 
with a quality of unity and a back pressure of 5 lbs. per square 
inch absolute. Clearance = 10%. 

If this engine delivers 27 H.P., and has a mechanical efficiency 
of 90%, what is the quality at the point of cut-off? 

11. In the previous problem, assume release to occur at 90% 
of the stroke with an absolute pressure of 30 lbs. per square inch. 
What is the quality at this point? 

12. A certain engine gives one horse-power hour at the shaft 
for every 20 lbs. of steam supplied. The steam has an initial 
pressure of 150 lbs. absolute and is dry and saturated when it 
arrives at the engine. The back pressure against which steam is 
exhausted is 4 lbs. absolute. 

(a) Find the thermal efficiency of this engine on the developed 
or shaft horse-power. 

(b) If the mechanical efficiency of the engine is 90%, what 
is the value of the thermal efficiency on the indicated horsepower? 




Volume 



CHAPTER IX 
COMPOUNDING 

71. Gain by Expansion. The cycle which gives a 
rectangular PF-diagram is the least economical of all the 
ideal cycles described in Chapter IV. This comes from 
the fact that none of the heat stored in the steam is con- 
verted into work when this cycle is used. Thus, if the 
cylinder shown by full lines in 
Fig. 83 operate on this cycle 
and be of such size that it will 
receive just one pound of steam 
per cycle, it makes available an 
amount of work represented by 
the area abed. The positive 
work done by the steam upon 
the piston is the equivalent of 
the external latent heat of 
vaporization while no use is 
made of the heat stored in the 
steam. This stored heat is re- 
moved as heat during the con- 
densation and exhaust, which give the lines be and cd. 

If a piece be added to the cylinder as indicated by 
the dotted lines, the same quantity of steam will make more 
heat available by expanding after cut-off, as shown by the 
curve be, the net work in this case being represented by 
the area abefd instead of by the smaller area abed. But 
the heat supplied is the same in both cases, namely that 
required to form one pound of steam at the pressure Pi, 
so that the use of a large cylinder and the incomplete ex- 

141 



j 




n 

i 

i i 
i i 

i ; 
i ! 



Fig. 83. 



142 



STEAM POWER 



pansion cycle results in the development of more work 
than can be obtained with the rectangular cycle from the 
same amount of heat. 

Obviously it would be theoretically advantageous to 
add still more to the length of the cylinder and allow the 
expansion to continue to back pressure, giving the com- 
plete expansion cycle as shown in Fig. 84, thus obtaining 
the maximum quantity of work at the expense of the heat 
stored in the steam supplied the cylinder. Practically, 
it is found inadvisable to continue the expansion to such a de- 




~--- ft 



Fig. 84. 



gree in reciprocating steam engines, because at low pressures 
the volume increases very rapidly for small pressure drops. 
Thus a great increase is necessary in the size of the cylinder 
if the last part of the expansion is to be completed, but 
the amount of work obtained is comparatively small, as 
shown by the small height of the long toe thus added to 
the diagram. This may result in an actual loss, because the 
increased friction losses of the very large cylinder may 
more than balance the small increase of net work gained 
by its use. It thus results that, in every real reciprocating 
engine, there is some point beyond which it is not economi- 
cal to carry the expansion, and the incomplete expansion 



a 


b' b' 


o 


(I :' 




c' 
c' 




V 


1 





COMPOUNDING 143 

cycle is therefore approximated in such engines rather 
than the cycle with complete expansion. 

Viewing the matter from another angle, a cylinder of 
a certain size may be assumed as shown in Fig. 85. The 
use of the rectangular cycle 
abed in this cylinder will make 
available the maximum quan- 
tity of work possible with the 
upper and lower pressures 
chosen. If cut-off be made 
to occur earlier as at b' , the 
expansion b'c' will result in a 
loss of the quantity of work 
obtained, as shown by the 
area b'bc' ', but the steam used Fig. 85. 

per horse-power will be less, 

so that there will be a gain in steam economy. Putting the 
cut-off still earlier will cause a still greater loss of work 
obtained from a cylinder of the chosen size, but theoreti- 
cally will result in greater economy of steam. 

Summing up, it may be said that the greater the ratio 
of expansion the greater should be the economy in the use 
of steam on a theoretical basis. 

The lower pressure is set in real engines by the pressure 
in the space into which the engine is to exhaust. If the 
engine is to be operated non-condensing, the atmospheric 
pressure determines the lowest possible exhaust pressure; 
if the engine is to be operated condensing, the exhaust 
pressure is set by the lowest pressure which can be eco- 
nomically maintained in the condenser. 

There is thus a real limit to the extent to which expan- 
sion can be carried in any real engine with a given initial 
pressure. A certain drop must exist at the end of the 
diagram, for reasons already explained, and an expan- 
sion line drawn backward from the top of the line repre- 
senting this drop will give the earliest possible cut-off 



144 STEAM POWER 

which can be used in the engine with a given initial 
pressure. 

The ratio of expansion can be further increased, how- 
ever, by raising the initial pressure as indicated by the 

dotted lines in Fig. 86, 

p| n anc l the limit in this direc- 

\ tion would come with the 

\ inability of materials of 

construction to withstand 
the resulting strains. 

These conclusions 
drawn from the facts 
V developed above must all 
p 86 be modified in the case of 

real engines, because of 
the effect of cylinder condensation. This has been shown 
to increase as the cut-off is made earlier and as the 
pressure (and therefore the temperature) range in a cyl- 
inder is increased. There is, therefore, a limit beyond 
which it is not advisable to carry the ratio of expansion in 
a single cylinder. 

Experience has shown that the best commercial results 
are obtained from simple engines, that is, those expanding 
the steam entirely in one cylinder, when (a) they are operated 
non-condensing, (b) the initial pressure is between 80 and 100 
pounds per square inch for the simpler forms of valves and up 
to 125 lbs. with the better forms of valves, and (c) the point 
of cut-off is at about | stroke with the simpler valves and at 
from I to J stroke with the better forms of valves. These 
values of cut-off correspond to nominal expansion ratios 
of 5 and 4 respectively and to lower values when clearance 
is taken into account. 

72. Compounding. If the ratio of expansion is to be 
increased above the values just given, some means must 
be used for the reduction of loss by condensation. This 
loss can be reduced by decreasing the surface exposed to 



COMPOUNDING 



145 




high-temperature steam and by decreasing the temperature 
range in a cylinder. Both of these results can be achieved 
by what is known as compounding. 

Assume that it is deemed advisable to produce a cycle 
similar to that shown in Fig. 87 (clearance neglected) and 
that in order to obtain 
high steam economy (low 
water rate) the ratio of 
expansion chosen is very 
much greater than four. 
No gain in economy 
would result from such 
excessive expansion in a 
single cylinder, in fact 
there would be a well- 
defined, unavoidable loss. 
But suppose that the high- 
pressure steam is admitted 
to a small cylinder such 
as that shown and is ex- 
panded to the point /, is 
then exhausted as shown 
by fg into the larger cyl- 
inder along gf and then expanded to the point c in the 
larger cylinder. The cycle produced is the same as that 
which would have been obtained by expanding entirely in 
one cylinder, but the surface of the clearance space of 
the high-pressure (H.P.) cylinder, which is exposed to high- 
pressure steam is smaller than it would be in a cylinder of 
the size required to hold the steam when fully expanded and, 
moreover, the lowest temperature to which it is subjected is 
that corresponding to the pressure at / instead of the 
much lower temperature corresponding to the pressure 
at d. 

The condensation which would occur in the H.P. cylinder 
would obviously be less than that which would result from 



Low Pressure Cylinder 
L.P. Cylinder 



Fig. 87. 



146 STEAM POWER 

the use of one large cylinder and, remembering that the 
greater part of the heat given up during condensation is 
received again by the steam during exhaust, it is obvious 
that approximately this same quantity of heat can again 
be given to the low-pressure cylinder walls. Thus, although 
there are two cylinders in which condensation and re- 
evaporation occur, and although the sum of the heat given 
to the walls of the high-pressure cylinder and that given 
to the walls of the low-pressure cylinder might be greater 
than that given to the walls of a single cylinder under similar 
conditions, the use of two cylinders results in a consider- 
able saving because loss in the high-pressure cylinder is 
practically wiped out by the exhaust of the heat concerned 
into the low-pressure cylinder. 

If the loss by radiation and conduction from the high- 
pressure cylinder be neglected, the result of the use of two 
cylinders is practically to limit the loss by condensation 
and re-evaporation to that occurring in the low-pressure 
cylinder. As the ratio of expansion in this cylinder is in 
the neighborhood of that common in simple engines, or 

even less, and as the tem- 
perature range is small, 
the net loss is also small. 
It is obvious that the 
smaller the surface of the 
high-pressure cylinder can 
be made, and the smaller 
the temperature range in 
a single cylinder, the 
smaller will be the net loss 
Fig. 88. by cylinder condensation 

• and re-evaporation. A 
saving should therefore be effected by using more than two 
cylinders, and it is not inconceivable that five or more might 
be used. The result of using five cylinders is shown in Fig. 
88, and it is evident that the clearance surfaces exposed to 




COMPOUNDING 147 

high temperatures, the temperature ranges per cylinder 
and the ratios of expansion per cylinder are all small. 
The gain in economy should therefore be correspondingly 
great. 

There are two limits to the possible multiplication of cylin- 
ders in this way. 

(1) As the number increases the radiating surface and 
therefore the heat lost by radiation increases. The extent 
of this effect can be appreciated by noting that every 
cylinder with the exception of the low-pressure cylinder is 
really an unnecessary addition, because the cycle could be 
produced entirely in the low-pressure cylinder. On the 
other hand, the surfaces of cylinders which operate at high 
temperature are small as compared with that which would 
be exposed to this temperature if the entire cycle were pro- 
duced in the low-pressure cylinder. 

(2) As the number of cylinders is increased, the first 
cost, the complexity and the cost of lubrication and attend- 
ance are all increased so that, for each installation, some 
number will be found beyond which the interest on the 
investment and the added cost of operation and mainte- 
nance would more than balance the saving of fuel. 

The second limit mentioned is the more important 
commercially, as it is the first one reached. For ordinary 
operating conditions in stationary power plants expansion 
in two cylinders generally gives the most economical results. 
The total ratio of expansion is generally between 7 and 16, 
that is, the volume of steam at release in the L.P. cylinder 
is from 7 to 16 times the volume at cut-off in the H.P. 
cylinder. For large pumping stations and large marine 
installations, expansion in three cylinders is generally 
considered the most economical, and total ratios of expansion 
of 20 or more are used. Four and five cylinders have 
been used, but the resultant gains do not seem to warrant 
any extensive installation of such units. 

Engines using more than one cylinder for the expansion 



148 



STEAM POWER 



of steam in the way just described are called multi-expan- 
sion engines, or compound engines, and the use of multi- 
expansion is spoken of as compounding. Custom has 
almost confined the use of the term compound engine 
to those in which only two cylinders are used in series as 
indicated in Fig. 89, and such engines are often spoken of 
as 2x engines. 

Engines in which three cylinders are used in series 
are called triple-expansion or 3x engines. With four and 
five cylinders in series the engines are known as quadruple 
or 4x and quintuple or 5x, respectively. 

In the case of triple-expansion engines of large size, 



H.P. Exhaust and L.P. Admission 




Boiler 



To 

Condenser 




Fig. 89. 



Fig. 90. 



the volume of the low-pressure cylinder required generally 
becomes so great that it is found economical to use two 
low-pressure cylinders instead of one. The flow of steam 
in such an engine is represented diagrammatically in Fig. 
90. This type is known as a four-cylinder, triple-expansion 
engine. 

All multi-expansion engines are generally operated 
condensing, and the choice of type is determined partly by 
the character of work to be done and partly by economical 
considerations. In all cases the boiler pressure must be 
chosen to suit the type of engine used. The pressures 
ordinarily used with the different types are given in 
Table V. 



COMPOUNDING 149 

TABLE V 
Boiler Pressure Commonly Used 



Type of Engine. 


Boiler Pressure. 
Pounds per Sq.in. Gauge. 


Simple 


80 to 125 


High-speed compound. . 


100 to 170 


Low-speed compound 


125 to 200 


Triple expansion and higher 


125 to 225 



73. The Compound Engine. The term compound 
engine will be used hereafter in the commercial way as 
referring to a 2x engine. Such engines may roughly be 
divided into two types, receiver and non-receiver engines. 
The latter are often called Woolf engines, after the man 
who first used this construction. 

A receiver engine has a vessel known as a receiver 
located between the two cylinders and so connected with 
them that the high-pressure, cylinder exhausts into the 
receiver and the low-pressure cylinder draws its steam 
from the receiver. By using a receiver the cylinders are 
made independent of each other so far as steam events 
are concerned; the high-pressure cylinder can exhaust 
at any time with reference to the events occurring in the 
low-pressure cylinder. 

A Woolf type has practically no receiver, the high-pres- 
sure cylinder exhausting directly into the low-pressure 
cylinder through the shortest convenient connecting pass- 
age. As the high-pressure cylinder must exhaust directly 
into the low-pressure cylinder it follows that cut-off must 
not occur in the latter until compression starts in the 
former; i.e., very near the end of the stroke. 

An engine with a receiver of infinite size would give 
a horizontal exhaust line for the high-pressure cylinder 
and a horizontal admission line for the low-pressure cylinder, 
since the small amount of steam given to or taken from the 



150 



STEAM POWER 



receiver would have no appreciable effect upon the pressure 
within that vessel. Neglecting throttling losses, the high- 
pressure and low-pressure cards would therefore fit together 
as originally indicated in Fig. 86. 

With receivers of finite size there are pressure changes 
during exhaust by the high- and admission to the low-pres- 
sure cylinders, and real valves and connections also cause 
certain throttling losses, so that the lines representing 
these events are not horizontal nor do they exactly coincide. 

A diagrammatic arrangement of the Woolf engine 
is given in Fig. 91 with idealized diagrams obtained by 




Fig. 91 



assuming hyperbolic expansions, no clearances, and no 
throttling losses. The pistons must make their strokes 
together in such engines, but they may move in the same 
direction, as shown in the figure, or in opposite directions. 

The ideal diagram would be that shown at (a) by the 
lines AbcdCDA. The idealized high-pressure diagram is 
abcda and the idealized low-pressure diagram is ABC DA. 
The exhaust line da of the high-pressure diagram and the 
admission line BC of the low-pressure diagram are pro- 
duced at the same time. Corresponding points on these 
two lines represent the common pressures assumed by the 
steam not yet exhausted from the high-pressure cylinder, 
the steam in the small connecting passage and the steam 



COMPOUNDING 



151 



already admitted to the low-pressure cylinder. As the 
movement of the low-pressure piston opens up volume 
faster than the high-pressure piston closes up volume, 
the volume occupied by the steam continues to increase as 
the low-pressure piston moves out, and its pressure there- 
fore decreases. 

The two diagrams are shown back to back at (b) in the 
figure and the horizontal line xX connects corresponding 




Fig. 92. 






points on the exhaust of the high pressure and the admission 
of the low pressure. 

Compound engines are also divided into two types 
on the basis of cylinder arrangement. When the axes of 
both cylinders coincide as shown in Fig. 92 they are called 
tandem compounds. When the axes are parallel as shown 
in Fig. 89, the engines are spoken of as cross-compound 
engines. 

74. Cylinder Ratios. The idealized diagrams of a com 
pound engine with infinite receiver 
volume are shown in Fig. 93 by 
abed and ABODE. The height of 
the high-pressure exhaust line is the 
same as that of the low-pressure 
admission line and represents the 
receiver pressure pr. The value of 
the receiver pressure is determined 
by the point chosen for cut-off in the low-pressure cylinder. 
Thus if cut-off in the low-pressure cylinder is made to occur 
earlier, as at some point c', the admission line for this 





_,c 


\ d 




b' 










B 

A 


V- 


c-sJ 


,D 


r 




— V L — 


^ E 



rr 



Fig. 93. 



152 STEAM POWER 

cylinder must move up to B'C and the receiver pressure 
must rise correspondingly. The exhaust pressure in the 
high-pressure cylinder would also rise an equal amount. 

Changing the point of cut-off in the low-pressure cylinder 
also produces another result. As the receiver pressure rises 
the work area of the high-pressure diagram is obviously de- 
creased, while that of the low-pressure diagram is increased. 
In a simple engine the area of the diagram becomes smaller 
the earlier the cut-off, and it should be noted that just the 
reverse of this occurs in the low-pressure cylinder of a com- 
pound engine. 

It is evident that the choice of the receiver pressure or 
of the point of cut-off in the low-pressure cylinder determines 
the relative areas of the high-pressure and low-pressure 
diagrams and it also determines the relative size of the two 
cylinders. The diagram of Fig*. 93 shows that late cut-off 
in the low-pressure cylinder calls for a larger high-pressure 
cylinder than does early cut-off. 

The ratio of the piston displacement of the low-pressure 
cylinder to that of the high-pressure cylinder is called the 
cylinder ratio. Designating this ratio by R, and using other 
symbols as in Fig. 93, 

R = j^ (61) 

y h 

The cylinder ratios chosen for real compound engines 
vary greatly in different designs and no given ratio has been 
proved the best for a given set of conditions. Normal 
practice gives the average values listed in Table VI, but 
cylinder ratios as high as 7 have been used with excellent 

results. 

TABLE VI 

Cylinder Ratios for Compound Engines 

Cylinder ratio 2£ 3i 4 4| 

Initial pressure (gauge) non-condensing 100 120 

Initial pressure (gauge) condensing 100 120 150 



COMPOUNDING 153 

The cylinder ratio to be used in a given case may be 
determined by any one of several considerations or by a 
combination of them, the latter being more often the case. 
Thus it may be deemed desirable to obtain the same amount 
of work from both cylinders; or to obtain equal temperature 
ranges; or to have cut-offs occur at the same fraction of 
the strokes; or to have the same total load on the two 
piston rods during admission; or to obtain the maximum 
possible uniformity of turning effort at the crank. The con- 
sideration of equal work is generally , , , 
regarded as the most important. 

75. Indicator Diagrams and Mean 
Pressures. The idealized diagrams 
for a compound engine with clearance, 
with incomplete expansion in both 
cylinders, and without compression 
are given in Fig. 94. The nominal 

total ratio of expansion would be L L -t-l H} but the total ratio 
of expansion taking account of clearance is 

Total ratio of expansion = ■= — , n , , . . (62) 

lH-\-tlH 

and the cylinder ratio is 

R = ^ (63) 

The mean effective pressures can be found from each 
of the diagrams in the ordinary way and the indicated 
horse-power of each cylinder determined therefrom. The 
indicated horse-power of the engine is then equal to the sum of 
the values obtained for the separate cylinders. 

It is often convenient to refer the mean effective pres- 
sure of all cylinders to the low-pressure cylinder as though 
this were the only cylinder acting. In the simple form 
of diagram, such as that shown in Fig. 93, it is obvious 
that this could be obtained by measuring the area AbcDEA, 



154 STEAM POWER 

dividing by the length AE and multiplying by the scale of 
the spring, just as though the diagram were all produced 
in one cylinder with the piston displacement equal to V L . 
In the case of the diagrams given in Fig. 94 a similar method 
could be adopted, or the mean effective pressure of each 
cylinder could be determined separately and then the 
equivalent pressure which would give the same result on 
the low-pressure piston could be determined analytically. 

Assume for this purpose that the mean effective pressure 
of the high-pressure is equal to p H pounds per square inch, 
that the mean effective pressure of the low-pressure cylinder 
is equal to p L and that the cylinder ratio is R. The strokes 
of all cylinders of a multi-expansion engine are generally 
equal, so that the piston areas are in the same ratio as the 
cylinder volumes (piston displacements). In the case of 
a 2x engine, therefore, the area of the low-pressure piston 
is R times as great as that of the high-pressure piston, 
and the pressure required on the low-pressure piston to do 
the same work as that done by pressure pn on the high- 

pressure piston will be -—. 
R 

In the case of a 2x engine therefore the total M.E.P. 

referred to the low-pressure cylinder is 

Vi^+Pl (64) 

This mean effective pressure acting on the low-pressure 
piston only would give the same indicated horse-power as 
is obtained with the two cylinders of the engine. 

In designing compound engines it is customary to 
determine the size of the low-pressure cylinder as though it 
were to do all the work expected of the engine by receiving 
steam at the highest pressure available and exhausting it 
at the lowest. The mean effective pressure which would 
thus be assumed to exist is the referred value pr just ex- 
plained. Having found the size of the low-pressure cylinder 



COMPOUNDING 155 

and the value of the referred M.E.P. the size of the high- 
pressure cylinder can be determined so that the work done 
by each cylinder will be just half of the total for which the 
engine is being designed. This size will have. to be such 
that the high-pressure mean effective pressure referred to 
the low-pressure cylinder (i.e., p H -i-R) is equal to half the 
total mean effective pressure referred to that cylinder. 
That is, the size will have to be so chosen that 

¥ i ^ 

ILLUSTRATIVE PROBLEM 

A double-acting compound engine is capable of developing 
500 I.h.p. The stroke is 18 ins.; revolutions per minute, 175; 
mean effective pressure referred to L.P. piston, 45 lbs. per square 
inch; cylinder ratio, 3^. Find cylinder diameters. 

From 

pLan 



I.h.p. 



«L.P. 



so that 



33,000' 

500X33,000 
45X1.5X175X2 

W 



/700 

\J85 



30 ins. (approx.), 

oO 



with the cylinder ratio equal to 3 J, 

«h.p. =---=200 sq.ins., 
3.5 

/200 

<W. = \/^ = 16 ms - (approx.). 

76. Combined Indicator Diagrams. When a compound 
engine is indicated, the diagrams of the two cylinders as 
drawn by the indicator are not directly comparable. The 
scales of pressure and volume are different on the two dia- 
gram's, and correction must be made for this fact before the 



156 



STEAM POWER 



diagrams can be compared. It is customary to do this and 
to draw the average high-pressure and low-pressure diagrams 
on the same set of coordinates in order to determine how 
well they approximate the ideal diagram that would be 
obtained in one cylinder operating between the extreme 
limits £>f pressure. 

Diagrams approximating those that would be obtained 
from high- and low-pressure cylinders are shown at (h) 
and (I) respectively, in Fig. 95, and the result of drawing 



b' s 




f . , 


. . 1 . . . i( 


"^3 

- 3 


I , . . . 

r — > 

| A-tmos. 


1 . . . .1 

U) 


V 




A.tmos. 


JUj^^ 




j^J 




Fig. 95. 



both to the same scales is shown at the left of this figure. 
The curves xn and xi show the variations of quality along the 
two expansion curves. 

Drawing the two diagrams to the same scales in this 
way is known as combining the diagrams and the result is 
known as a combined diagram. 

The curves SS and S'S' added to the combined diagram 
are saturation curves. They do not, in general, form a 
continuous curve, because of the different quantities of 
steam contained in the two clearances and because any 



COMPOUNDING 



157 



moisture in the high-pressure exhaust is generally removed 
in the receiver. The volumes occupied by clearance steam 
at initial pressures are indicated by the points V and B' 
respectively. The lengths b'S and B'S' approximately 
represent the volumes that would be occupied by cylinder 
feed when in each cylinder if dry and saturated. 

A combined diagram for a triple-expansion engine is 
shown in Fig. 96. The heavy lines give diagrams con- 
structed so as to represent 
as nearly as possible what 
may be expected to occur in 
the cylinders of such an en- 
gine, assuming perfect valve 
action and hyperbolic expan- 
sions and compressions. The 
dotted diagrams indicate the 
shapes that would be drawn 
by indicators applied to the 
real cylinders. The numer- 
ous sharp angles are due to 
overlapping of events, one 
cylinder suddenly starting to draw from a receiver while 
another is exhausting. It will be observed that the dotted 
diagrams do not contain any of these sharp angles, but 
that their general outline forms a fair average of them. 

The curve cd is a rectangular hyperbola drawn as a 
continuation of the assumed hyperbolic expansion line of 
the high-pressure cylinder. The failure of the expansion 
lines of the other cylinders to fall upon this curve is ex- 
plained by quality changes, different quantities of clearance 
steam in the different cylinders and withdrawal of moist- 
ure from steam exhausted to receiver before admission to 
the following cylinder, 




Fig. 96. 



158 STEAM POWER 



PROBLEMS 



1. Find the size of the cylinders of a double-acting compound 
engine, which is to give 600 I.h.p., when using steam at a pressure 
of 150 lbs. per square inch absolute, and having a back pressure 
of 2 lbs. per square inch absolute. The cylinder ratio is to be 4, 
and the total ratio of expansion 12, piston speed 750 ft. per minute, 
and R.P.M. =150; diagram factor is 80%. 

2. Given a 200 H.P. compound Corliss engine with cut-off 
in the H.P. cylinder at 60% stroke. Ratio of expansion is 7; 
clearance is 7%; card factor is 70%; pressure at the H.P. cyl- 
inder is 165 lbs. absolute. Find 

(a) Cylinder ratio; 

(b) Theoretical and actual M.E.P.; 

(c) Determine size of four engines, and select the best one. 

\-\-% CI 

Note - K = %(C .o.)+%ci Xcyl - ratio - 

3. Given a compound engine 18X40 ins., having a stroke of 
28 ins. Steam pressure is 165 lbs. per square inch absolute; 
cut-off in H.P. cylinder occurs at 62% stroke; clearance equals 
16%; back pressure equals 5 lbs.; R.P.M. equal 150. Find 

(a) Cylinder ratio; 

(b) Ratio of expansion; 

(c) Actual M.E.P.; 
{d) I.h.p. 



CHAPTER X 
THE D-SLIDE VALVE 

77. Description and Method of Operation. The simple 
D-slide valve, shown in place in Fig. 97, is so named because 
of the similarity of its section to the letter D. It is located 
in the steam chest, rides back and forth upon its seat and 




Fig. 97. 



serves to connect the two ports alternately with steam 
and exhaust spaces respectively in order to give the neces- 
sary distribution of steam. 

The valve has to perform the following functions for 
each end of the cylinder during each revolution of the 
engine : 

(1) It connects the proper port to the steam space or 

159 



160 STEAM POWER 

steam chest at such a time that steam can enter the cylinder 
as the piston moves away from the head. 

(2) It shuts off this port and thus cuts off the supply 
of steam when the piston has completed a certain definite 
fraction of the stroke. 

(3) It connects the port with the exhaust cavity shortly 
before the piston reaches the end of the stroke, thus effecting 
" exhaust " or " release "; and 

(4) It shuts off the port again when the piston has com- 
pleted the proper fraction of the next stroke, thus trapping 
in the cylinder the steam which is compressed during 
the remainder of the stroke. 

Engine Crank ^ 
Main Connecting 




Fig. 98. 



It is obvious that the valve must be reciprocated upon 
its seat and that its motion must be connected with that of 
the piston in some way so that the proper phase relation 
may be retained. This could be effected by the system 
shown diagrammatically in Fig. 98, a small crank operating 
on the end of a connecting rod giving the valve its short 
stroke just as the main crank fixes the longer stroke of 
the piston. Such an arrangement would, however, be 
very inconvenient with many real engines, as the valve would 
be located too far from the center line of the cylinder. 

It is customary to use what is known as an eccentric 
for the purpose of operating the slide valve. The parts 
and arrangement of an eccentric, together with an illus- 
tration of the way in which it is mounted on the shaft of 




Fig. 99. — Parts of Eccentric. 



161 



162 



STEAM POWER 




THE D-SLIDE VALVE 



163 



an engine are shown in Figs. 99, 100 and 101. The motion 
it gives the valve is exactly the same as that imparted 
by the crank first assumed, and it can 
easily be shown that it is the exact 
equivalent of such a crank. 

Assume, for example, a crank such 
as that shown in Fig. 98 with a length 
of arm or throw equal to a. If the 
crank pin is made larger while other 
parts of the crank remain the same, as 
shown in Fig. 102, the crank mech- 
anism is not essentially altered; the mo- 
tion which it would impart to a connecting rod is not 
changed. If this process of enlarging the pin be continued 




Fig. 101. — Eccentric 
on Vertical Engine. 







Fig. 102. — Equivalence of Crank and Eccentric. 



until the pin has become large enough to surround the 
shaft and if the crank arm be then removed so that what 
was the crank pin is fastened directly on the shaft, an 






164 



STEAM POWER 




Fig. 103. 



Fig. 104. — Slide Valve without Lap. 



TAve Steam Space 

Exhaust CavityConnected 
to Exhaust Eipe 




Fig. 105. 



THE D-SLIDE VALVE 



165 



eccentric results. It is the exact equivalent of the original 
crank; its center, which is the center of the crank pin, 
revolves about the center line of the shaft in a circle with 
a radius a just as in the original mechanism. 

The eccentric makes it possible to place a short crank 
(short arm) upon a large diameter shaft without having to 
cut the shaft away as shown in Fig. 103, and it is therefore 
very useful for driving valves. 

78. Steam Lap. The simplest possible form of D-slide 
valve would just reach the outer edges of the ports when 
in its central position as shown in Fig. 104. The crank 
driving it (that is the crank equivalent to the eccentric 
which would probably be used in a real case) would have 
to be located 90° ahead of the engine crank in the direction 
of rotation, as can easily be seen by consulting Fig. 105, 
which illustrates the mechanism in various critical positions. 
The illustration shows that such a valve would give full 
stroke admission, thus producing a rectangular cycle which 
has already been shown to be very inefficient as a means of 
obtaining work from the heat used in 
forming steam. 

If cut-off is to occur before the 
end of the stroke, the edge of the 
valve must return and close the port 
before the piston reaches the end of its 
stroke. But since the crank mechan- 
ism does not permit the valve to 
remain stationary in any one posi- 
tion, such early cut-off could only 
over-traveled, as shown in Fig. 106, 



Steara. Space 

-« — Valve Travel 

r , ' -~T"7 







Piston Travel 



Fig. 106. 



occur if the valve 
and this would un- 
fortunately result in connecting the working end of the 
cylinder to exhaust and in admitting steam to the other 
side of the piston at such a time as to oppose the piston's 
motion. The solution of the difficulty lies in making 
the valve longer, so that when in its central position it 
overlaps the outer edges of the ports as shown in Fig. 107. 



166 



STEAM POWER 



The amount of overlap of the outer edge is called the out- 
side lap, and when steam is admitted by the outer edges of 
the valve, as in the case under discussion, it is also called 
the steam lap. 

With such an arrangement the valve must be drawn 
out of its central position by the amount of the lap when 
the piston is at the end of its stroke as shown by a in Fig. 







3" 



Fig. 107. — Steam and 
Exhaust Lap. 




Fig. 108. — Lap a and Lap 
Angle a. 



108 in order that steam may be admitted just as the 
piston starts to move. It follows that the crank driving 
the valve must be more than 90° ahead of the engine 
crank and that it must be ahead by the angle required 
to move the valve ' a distance equal to the outside lap. 
This angle, represented in the figure by a, is called the lap 
angle. 

79. Lead. In real engines it is further desirable to start 
the admission of steam just before the piston arrives at 
the end of its stroke. This assists in bringing the moving 
parts to rest, raises the pressure in the clearance to full 
value before the piston starts, and gives a wider opening 
through which the steam can flow during the early part of 
the stroke, thus reducing wiredrawing and loss of area at 
the top of the diagram. If the valve is to open before the 
piston reaches the end of its stroke, the crank driving it 
must be shifted still further ahead of the engine crank. 
It must be shifted ahead by an angle which will draw 
the valve through the distance which will give the desired 
opening of valve with the piston at the end of its stroke 
as shown by b in Fig. 109. The angle required, indicated 



THE D-SLIDE VALVE 



167 



by /3, is known as the angle of lead, and the width 
of the steam opening with engine crank on dead center, 
i.e., the distance b, is known as the lead. The lead 
varies from less than yg in. on small engines and with low 
speeds up to over § in. on large engines and with very high 
speeds. 

80. Angle of Advance. The eccentric or valve-operating 
crank must be ahead of the engine crank by an angle equal 
to 90°+ angle of lap a + angle of lead /3, as can be seen 




Fig. 109.— Lead b and Lead Angle /3. 



by an inspection of Fig. 109. The sum of a and (3 is called 
the angle of advance and will be represented by 8. This 
is the number of degrees in excess of 90 by which the eccen- 
tric leads the engine crank. 

Fig. 109 shows that cut-off in an engine fitted with 
a valve having lap and lead must occur when the engine 
crank has turned through an angle equal to 180 — 2a, because 
the valve will then have returned to the closed position. 
Apparently, cut-off can be made to occur at any point 
in the stroke by properly choosing the value of a, but it 
will be discovered later that the exhaust events set a limit 
to increase in the value of this angle and hence do not per- 
mit of cut-off occurring earlier than a certain fraction of the 
stroke. 

81. Exhaust Lap. Inspection of Fig. 105 will show that 
the simple valve without lap originally assumed will give 
no compression, because the cylinder end is connected to 
the exhaust cavity for the entire stroke. Inspection of all 



168 STEAM POWER 

the changes which have been suggested in the subsequent 
paragraphs will show further that if the inner edges of the 
valve are left in the original positions the exhaust events 
will be considerably distorted in the case of a valve having 
steam lap and lead. 

This trouble may be remedied by moving the inner 
edges of the valve closer together, making the exhaust 
cavity in the valve shorter and giving inside lap as shown 
in Fig. 107 by b. When the inner edges of the valve control 
exhaust, as in the case of the valve under discussion, this 
inside lap is also called exhaust lap. 

The length of the valve, the lap and the lead are gen- 
erally chosen so as to give the desired arrangement of 
admission and cut-off and then the exhaust edges are so 
located as to give desirable release and 
compression. In some forms this necessi- 



fi/>;, J/..*-$. ss zrs tates the use of an exhaust cavity in the 

Fig. 110. valve such as that shown in Fig. 110. 

The amount by which the edges of the 

valve fail to meet the inner edges of the port is spoken of as 

negative inside lap. This dimension is indicated by c in the 

figure. 

It should be noted particularly that all measurements 
of lap are made with the valve central on its seat and 
that the measurement of lead is made with the piston at 
the end of its stroke, i.e., with the engine crank on dead 
center. 

82. The Bilgram Diagram. The action of all slide 
valves could be studied by means of drawings of the actual 
mechanism, as has been done in preceding paragraphs, but 
such a method is time and space consuming. Numerous 
diagrams such as the Elliptical, the Sweet, the Zeuner and 
the Bilgram have been developed for the purpose of simpli- 
fying and expediting such studies and, when properly 
understood, they are very convenient. The scope of this 
book does not permit a discussion of all of these diagrams 



THE D-SLIDE VALVE 



169 



and, since the Bilgram diagram is probably the most gener- 
ally applicable, attention will be confined to it. 

The construction of this diagram is illustrated in Fig. 
111. The point represents the center of the engine 
crank shaft and the two circles drawn about this point 
as a center represent respectively the paths traveled by the 




Fig. Ill, 



pin of the valve crank and the pin of the engine crank. 
These circles are drawn to any convenient scales. 

The diagram is conventionally drawn in such a way 
that the line OM represents the head end dead center 
position of the crank and in all subsequent paragraphs the 
relative positions shown by the small sketch in Fig. Ill 
will be assumed. The cylinder will be assumed to the 



170 STEAM POWER 

left of the shaft and the engine will be assumed to run 
" over." 

With the crank in position OM, the eccentric (equivalent 
crank) must be in the position OB, ahead of the crank by 
an angle 90°+g:+/3 = 90+5. The valve must then be 
displaced to the right of its central position by an amount 
represented by the distance DB, if a small correction for 
" angularity " of the valve connecting rod be neglected. 
As rotation continues, horizontal distances corresponding 
to this line will always give the instantaneous valve dis- 
placements. For position OB', for instance, the valve dis- 
placement will be D'B' . 

If the angle 8 is now laid off above OX, locating the point 
Q as shown, a perpendicular QE dropped upon OX from this 
point will equal in length the line DB, and will therefore 
show the valve displacement when the crank is in head end 
dead center position OM. This must be true, because the 
triangles QOE and BOD are similar and have the sides 
OQ and OB equal to the radius of the same circle. 

The perpendicular QE is really a perpendicular dropped 
upon the extension of the line representing the crank posi- 
tion, and it is a general property of this diagram that a line 
starting at Q and perpendicular to the line representing 
any chosen crank position (or an extension of that line) 
will show by its length the displacement of the valve when 
the crank is in the chosen position. Thus assume the engine 
crank to rotate through the angle 7 to the position OM' . 
The eccentric will have rotated to B' and the valve dis- 
placement will be represented by D'B' . A perpendicular 
drawn from Q upon OX', the extension of the crank posi- 
tion, gives QE equal to B'D' and hence representing the 
valve displacement to the same scale. 

This construction drawn for different crank positions 
OA, OM, OMi, OM 2 , etc., is shown in Fig. 112, the dash- 
dot radial lines about Q representing the various values of 
the valve displacement. The number of each of these 



THE D-SLIDE VALVE 



171 



lines indicates the crank position to which it corresponds. 
It will be seen that the displacement increases in value 
until the crank position 0M% is reached, after which it 
decreases again. 



Steam Lap 
Circle' 



Steam Lead 




Fig. 112. 



Since the opening 
ment minus the lap, as 
by which the valve is 
found by subtracting 
placement the amount 
head end dead-center 
to lap plus lead, and is 



to steam is equal to the displace- 
shown in Fig. 109, the actual amount 
open for any crank position can be 
from the corresponding valve dis- 
of lap possessed by the valve. For 
position, the displacement is equal 
shown by QE in Fig. 112. Subtract- 



172 STEAM POWER 

ing the lead EF, the remainder FQ gives the lap of the valve. 
A circle drawn about Q with radius equal to QF (or a circle 
drawn about Q and tangent to the line L) will cut off of 
the lines representing valve displacement the amount 
representing the part of each displacement used in over- 
running the lap of the valve. The remainders, that is the 
parts of the lines radiating from Q in Fig. 112 which are 
outside of the lap circle, must then represent the amounts 
by which the valve port is actually open. 

It will be observed that the valve is open by the amount 
of the lead when the crank is on dead center, position OM. 
The crank position for which the valve displacement is just 
equal to the lap, and hence at which the valve is just begin- 
ning to open, can be found by drawing a tangent through 
to the lower side of the lap circle and then extending 
it to give the crank position OA in Fig. 112. 

As the crank rotates clockwise from this position, the 
valve opens wider until, when position 0M% is reached, 
the greatest valve opening exists. Further rotation results 
in partial closure of the valve and, when the crank has 
finally rotated into position OC, the valve has just closed, 
that is, cut-off has occurred, the displacement being just 
equal to QG, the steam lap. 

Thus this diagram, as so far developed, indicates crank 
positions for admission and cut-off and the values of valve 
displacement and valve openings for all intermediate 
crank positions. 

ILLUSTRATIVE PROBLEM 

A certain valve has an external steam lap equal to 1| ins. 
The lead is ^ in. and the throw of the eccentric is 2\ ins. (a) Con- 
struct such parts of the Bilgram diagram as are necessary to 
indicate "head end" crank positions for admission, maximum 
valve opening and cut-off. (b) Indicate on this diagram the 
amount of valve opening at various crank positions between 
admission and cut-off. (c) Determine the value of the angle of 
advance. 



THE D-SLEDE VALVE 



173 



Draw a circle with radius equal to the eccentric throw, 2§ 
ins., using any convenient scale. This circle is designated by abed 
in Fig. 113. Draw about the same center another circle of any 
convenient size. Draw in the horizontal diameter ac and extend 
as shown. On the right-hand side of the circle draw the line ef, 




Fig. 113. 



parallel to the horizontal axis and a distance above it equal to the 
lead, ye hi., to the same scale as that chosen for eccentric circle. 
The steam lap circle must have its center Q on the upper right- 
hand quadrant of the eccentric circle, and it must be tangent 
to the line ef. Its radius must equal the steam lap, l£ in. to scale. 
Therefore, with compass points set the proper distance apart, find 
the center Q, about which a 1^-in. radius circle will just be tangent 
to the line ef, and draw the steam lap circle. 



174 



STEAM POWER 



The crank position at admission is found by drawing the line 
iO so that, if extended, it is tangent to the lower side of the 
steam lap circle. 

The crank position at cut-off is found by drawing the line 




J^Ineide lap 




Fig. 114. 



M""0 in such position that it is tangent to the upper part of 
the steam lap circle. 

The crank position for maximum valve opening is found by 
drawing the line M"0 in such position that a line through QO 
will be perpendicular to it. The amount of valve opening at this 



THE D-SLIDE VALVE 175 

crank position is shown by the length of the part of this per- 
pendicular line outside of the steam lap circle, i.e., the distance 
Og interpreted according to the scale chosen for eccentric and 
steam lap circles. 

When the crank is in position M'O, the length of hi, interpreted 
to scale, gives the amount by which the valve is open to steam. 

When the crank is in position M'"0, the length of jk, inter- 
preted to scale, gives the amount by which the valve is open to steam. 

The angle indicated by 5 is equal to the angle of advance 
because of the property upon which the construction of this 
diagram is based. 

83. Exhaust and Compression. The exhaust edge events 
can be shown on the Bilgram diagram by a method similar 
to that used for the steam edge events. The direction in 
which valve displacements occur are indicated in the upper 
part of Fig. 114 in which the crank and eccentric circles 
have been drawn to such scales that they coincide. In- 
spection of the small sketch in the lower part of the figure 
will show that head end release must occur when the valve 
has traveled a distance equal to the inside lap to the left 
of its central position. A crank position OR drawn tangent 
to the lower part of a circle about Q with radius equal to 
the inside lap will, therefore, be the crank position at re- 
lease. Clockwise rotation from this position will result 
in a wider opening to exhaust until position 0M\ is reached, 
after which the valve will begin to close. Final closure 
will occur when the crank reaches position OK, the exten- 
sion of which is tangent to the top of the exhaust lap circle. 
At that time the valve will have returned (moving from left 
to right) and will still have to move a distance equal to 
the exhaust lap before attaining a central position. 

ILLUSTRATIVE PROBLEM 

Given the «exhaust lap of a D-slide valve equal to f in. ; the 
steam lap If ins.; the throw of the eccentric, 2 ins.; and the 
lead | in. Find the angle of advance, the maximum port opening 
to steam and to exhaust, and the crank positions of cut-off, release, 
compression and admission for the head-end of the cvlinder. 



176 



STEAM POWER 



Draw the eccentric (and crank) circle with a radius equal 
to 2 ins., and draw the horizontal diameter as in Fig. 115. 

Draw a horizontal line in the upper right-hand quadrant at a 
distance of |+1| ms - above the horizontal diameter. Locate 
the point Q at intersection. 




Fig. 115. 



Draw the steam lap circle with a radius 1^ in. and the exhaust 
lap circle with a radius f in. 

The angle of advance is the angle between OQ and the hori- 
zontal. 

The maximum opening to steam is given by the distance 
Oa = f in. The maximum opening to exhaust is given by the 
distance 06 = If in. 

The crank positions shown are obtained by drawing lines 



THE D-SLIDE VALVE 



177 



tangent to the lap circles. A represents admission; C, cut-off; 
R, release, and K, beginning of compression. 

The piston positions at the times of these events are given 
to reduced scale by vertical projection. 

84. Diagram for Both Cylinder Ends. The complete 
diagram for the head end cylinder is shown in Fig. 114 with 
all critical crank positions marked. The positions for the 
crank end of the cylinder can be found in a similar way by 
constructing a diagram in which the point Q and the lap 
circles are located in the opposite quadrant. The resulting 




Fig. 116. 



diagram for both cylinder ends, with laps the same for both 
ends of the valve, is given in Fig. 116. 

85. Piston Positions. The valve events might be studied 
entirely in conjunction with crank-pin positions, but it is 
more convenient and customary to consider them in connec- 
tion with piston positions. Piston positions corresponding 
to different crank-pin positions could be found by drawing 
the mechanism to scale for each different position as shown 
in Fig. 117 for piston positions 1 and 2. 

It is obvious that this would involve a great deal of work 
and that, if drawn to large scale, it would consume a great 



178 



STEAM POWER 




deal of space. Further, it is 
convenient to be able to locate 
relative piston positions on the 
line which serves as the hori- 
zontal diameter of the crank 
circle of the Bilgram diagram. 

The method used depends 
upon the fact that the motion 
of the crosshead is exactly the 
same as that of the p'ston, 
so that if the motion of the 
crosshead end of the connect- 
ing rod can be followed, it 
will be equivalent to following 
the motion of the piston itself. 
It should also be noted that 
the diameter of the crank cir- 
cle must be equal to the 
stroke of the engine. 

Assume now, that the point 
b in Fig. 117 be taken to rep- 
resent the position of the pis- 
ton when it is really in posi- 
tion 1. When the piston has 
moved to position 2, the cross- 
head will have moved from a 
to a' and the crank pin from 
b to b' . If with a' as a center 
the connecting rod be swung 
down to the horizontal its 
right-hand end will arrive at 
the point c. The distance be 
must then represent the dis- 
tance that crosshead (and pis- 
ton) have moved from dead- 
center position because a b and 



THE D-SLIDE VALVE 



179 



a'c both represent the length of the connecting rod and c 
must therefore be as far to the right of b as a' is to the right 
of a. The point c may therefore be taken to represent 
piston position when the connecting rod is in the position 
a'V. 

In general, if the horizontal diameter of the crank 
shaft be taken to represent the stroke of the engine, the pis- 
ton position corresponding to any crank position can be 
found by taking a radius equal to the connecting-rod length 
(to the same scale as the circle) and striking an arc from the 




Fig. lis. 



crank-pin position, using a center on the horizontal line on 
the cylinder side of the crank circle. 

An approximate method is also used for finding the piston 
position. Instead of projecting down from the crank-pin 
position with an arc, such as b'c in Fig. 117, a vertical line 
through the crank-pin position is used. Such a line would 
give c' as the piston position when c is really correct. This 
method would give accurate results with a connecting rod 
of infinite length. For ordinary lengths of rod, however, 
the results are far from correct. The error is said to be due 
to the angularity of the connecting rod. 

The effect of the angularity of the connecting rod is 
shown in Fig. 118 for different positions. On the outstroke 
the piston is always farther ahead than the rectilinear pro- 



180 



STEAM POWER 



jection would indicate. On the return stroke the piston is 
always behind the position indicated by rectilinear projection. 




Fig. 119. 



86. Indicator Diagram from Bilgram Diagram. Since 
the piston positions corresponding to different crank posi- 



THE D-SLIDE VALVE 181 

tions can be determined, it is a comparatively simple matter 
to construct the indicator diagram which theoretically 
would be given by an engine fitted with a valve of certain 
dimensions. It is necessary to assume the upper and lower 
pressure and also to assume the form of the expansion 
and compression curves. These are generally taken as 
rectangular hyperbolas. 

The method of constructing an indicator diagram from 
the Bilgram diagram is shown in Fig. 119. The crank- 
pin positions for admission (A), cut-off (C), release (R) 
and beginning of compression (K) are first found. These 
pin positions are then projected to the horizontal diameter 
by means of arcs with radius equal to the connecting-rod 
length and with centers on the line MM produced to the 
left. The intersections a, c, r and k indicate the piston 
positions at which the corresponding events occur. These 
are then projected vertically downward to intersect the 
proper pressure lines and the card is drawn through the 
intersections. 

Diagrams constructed in the same way, but for both 
head and crank ends, are given in Fig. 120. A symmetrical 
valve was assumed, that is, one built exactly alike on head 
and crank ends. • The diagrams show that such a valve 
cannot give the same results for both cylinder ends because 
of the effect of the angularity of the connecting rod. It 
is most evident in the case of cut-off. The cut-off in this 
case occurs just before three-quarter stroke for the head end 
and just after half stroke for the crank end of the cylinder. 
All other events are distorted in the same way, but the actual 
lengths of the variations are not as great as in the case of 
the cut-offs and therefore the distortion is not as obvious. 

The effect of the angularity of the connecting rod upon 
the diagrams can be remembered easily if it is noted that all 
valve events occur later with respect to piston position on 
the outstroke and earlier on the instroke than they would 
with a connecting rod of infinite length, 



182 



STEAM POWER 



It is possible to " equalize " the cut-offs, that is, make 
them occur at the same fraction of the stroke by using 
unequal steam laps at opposite ends of the valve, but this 
will result in still further distortion of admissions, as can be 
seen by constructing a Bilgram diagram for this case. 
Similarly, the compressions can be equalized by the use of 



Conrpression 



1T.E. Admission 




C.E. Admission 



Fig. 120. 



unequal exhaust laps, but this results in distortion of the 
release events. 

Various linkages have been developed which are so 
arranged that they distort the motion of the valve to just 
the extent necessary to counterbalance the effects of the 
angularity of the connecting rod. The scope of this book 
does not, however, permit a discussion of such valve 
gears. 



THE D-SLIDE VALVE 183 

87. Limitations of the D-slide Valve. The simple 
valve discussed in the preceding paragraphs has numerous 
limitations and is therefore only used on small and cheap 
engines, or in cases where economy in the use of steam is 
not essential. This valve, when used with steam entering 
over the outside edges as previously considered, is pressed 
to its seat by the live steam acting over its entire upper 
surface. This pressure is practically unbalanced, as the 
greater part of the lower surface of the valve is subjected 
to the low pressure of the steam being exhausted. As a 
result the friction to be overcome in moving the valve is 
very great and there is an appreciable loss from this source. 

Further, the shape of the valve makes necessary the use 
of long ports which form part of the cylinder clearance 
and which are alternately exposed to live and to exhaust 
steam with results previously discussed. These ports can 
be decreased in length by increasing the length of the valve, 
but this in turn increases the area exposed to high pressure 
and hence increases the friction loss. 

It can be shown by means of the Bilgram diagram 
that, if a cut-off earlier than about f stroke is desired, 
the angle of advance, the amount of steam lap and the size 
of the eccentric must all be made very great. This results 
not only in large friction losses, but also in very early release 
and compression, because of the great angle of advance. 
As a result, slide valves of the simple D type are seldom used 
when a cut-off earlier than J to ]| stroke is desired. It 
should be remembered in this connection that the simple 
engine generally gives its best economy with a cut-off of 
about \ stroke. 

The drawing of lines representing the opening of the 
valve to steam as in Fig. 112 will show that this simple 
valve is further handicapped by the very slow opening 
and closing of the steam ports, causing a great amount of 
wire drawing with a corresponding loss of diagram area. 
In order to get an adequate opening to steam the valve 



184 



STEAM POWER 



must also be given a great displacement and, since this 

occurs under great pressure, it results in great friction loss. 

The unbalanced feature can practically be overcome 

by rolling up the valve 
.sieam spaces^^ and ports about an axis 

parallel to the length of 
the cylinder. This gives 
what is known as a pis- 
ton valve, shown dia- 
grammatically in Fig. 
121 

It can also be par- 
tially overcome by using 
a balance plate or ring of some kind between the top of 
the valve and the inside of the steam-chest cover, so 
arranged that live steam is excluded from the greater part 
of the upper surface of the valve. Valves of this type 
are generally called balanced slide valves and are used on 
many high- and medium-speed engines. 

The valve travel required for obtaining a given opening 




■Steam Ports 



Fig. 121.— Piston Valve. 




T 

w/ymm 

Fig. 122.— Allen Double Ported Valve. 



can be decreased and the rate of opening and closing can be 
increased by the use of multiported constructions. These 
are so arranged that two or more ports open or close at the 
same time, so that the total movement required for a given 
opening is divided by the number of ports and the rate of 
opening and closing is multiplied in the same proportion. 
One simple type of double-ported valve is illustrated in 
Fig. 122. 

When several ports are used the valve often becomes 



THE D-SLIDE VALVE 185 

a rectangular frame crossed by a number of bars and is 
known as a gridiron valve, because of its appearance. Such 
valves are often combined with balance plates and give 
very satisfactory results. 

A number of designs of slide valves have been developed 
for the purpose of making cut-off independent of the other 
events. Many of these use a separate cut-off valve which 
either controls the steam supply to the main valve or else 
rides on the main valve and controls cut-off by covering 
ports in that valve. Devices of the latter type are called 
riding cut-off valves. They are either driven by separate 
eccentrics, or by linkage from the eccentric controlling 
the main valve, the linkage being so arranged as to give 
the proper relative motion between main and auxiliary 
valves. In such designs the main valve is proportioned 
so as to give the desired admission, release and compres- 
sion and the cut-off is then taken care of by proper adjust- 
ment of the cut-off valve. 

88. Reversing Engines. It was shown in one of the early 
paragraphs of this chapter that the eccentric must be set 
90°+ angle of advance ahead of the crank, ahead meaning 
in the direction of rotation. To cause the engine to revolve 
in the opposite direction, that is, to " reverse " the engine, 
it is therefore only necessary to shift the relative positions 
of eccentric and crank so that the eccentric leads the crank 
by 90°+ 8 in the new direction of rotation. This corre- 
sponds to shifting ahead (in first direction of rotation) 
through an angle equal to 180 — 25 or shifting backward 
through an angle equal to 180+26, as can be seen by inspec- 
tion of Fig. 109. 

In practice it is generally more convenient to use two 
eccentrics, one set properly for rotation in one direction 
and the other set properly for rotation in the opposite direc- 
tion. This arrangement is shown diagrammatically in 
Fig. 123. This figure is drawn for a vertical engine and in 
such position that the engine is on crank-end dead center. 



186 



STEAM POWER 



The point P represents the position of the center of the crank 
pin; the point / represents the position of the equivalent 
crank (center of eccentric) which 
drives the valve for " forward," 
" ahead " or clockwise rotation; and 
the point b represents the position of 
the equivalent crank which drives the 
valve for " backing," " reverse," or 
counter-clockwise rotation. 

The real mechanism, in one of its 
numerous forms known as the Stephen- 
son Link Gear, is shown in perspective in Fig. 124. The 
forward eccentric corresponds to / of Fig. 123 and the 
backing eccentric corresponds 




Reversing or 
Suspension.Arm 

<Backing^ 
.Position 



""Reverse or „ 
Weight Shaft 

Valve Stem 




to b of that figure. The 
eccentric rods are fastened 
to opposite ends of a curved 
" link " and move the valve 
through a " link block " 
fastened to the end of the 
valve stem. In the position 
shown in the figure the link 
is in such position that the 
forward eccentric operates 
practically directly on the 
valve stem so that the valve 
motion is practically entirely 
governed by that eccentric. 
If the reverse shaft were to 
be rotated clockwise into the 
backing position, the "sus-. 
pension rods " would pull 
the link over until the eccen- 
tric rod of the backing eccentric was directly under the 
valve stem. Under such conditions the valve motion 
would be controlled almost entirely by the backing 



Backing 
" Ecc, 



Fig. 124. — Stephenson Link Gear. 



THE D-SLIDE VALVE 187 

eccentric and the engine shaft would rotate counter-clock- 
wise. 

If the mechanism were so set that the link block occupied 
a position on the link between the ends of the two eccentric 
rods, the valve motion would be controlled by both eccentrics 
and would be a compromise between the motions given by 
either eccentric separately. It is characteristic of this gear 
that the cut-off is latest when either one or the other eccentric 
is fully " in gear " and that it becomes earlier as the link 
block approaches the center of the link. With the link 
block in the center of the link the valve does not open at 
all, i.e., the cut-off occurs at zero stroke. 

There are numerous other forms of link gears, the best 
known being the Gooch, the Allan and the Porter- Allen. 
There are also numerous reversing mechanisms known as 
radial gears in which the motion of the valve is controlled 
by means of a " radius rod " which can be set to give the 
desired valve motion. The valve motion is obtained in- 
directly through the radius rod from an eccentric, from the 
crank, or from the connecting rod. The limits of this 
book do not permit a detailed discussion of these forms. 

89. Valve Setting. From what has preceded it will 
be evident that it is not only necessary that a valve and 
its seat and driving mechanism be correctly designed, but 
also that the various parts must be correctly connected up 
in order that the valve may move in its proper phase rela- 
tion with respect to the piston. 

Adjusting the mechanism in such a way that the proper 
phase relations are obtained is known as setting the valve. 
This can be done with fair accuracy by a simple study of 
the mechanism in various positions, as will be shown below, 
but it is always advisable to check the setting by means of 
indicator diagrams taken after the setting is completed. 
Such diagrams will often show errors of such character or 
size that they cannot be determined by measurement on 
an engine which is not operating. 



188 



STEAM POWER 



Before beginning operations it is always advisable to go 
over the entire engine carefully and to eliminate excessive 
lost motion at all pins and bearings in order that the relative 
positions of parts obtained while setting the valve may 
approximate those which will be obtained when the engine 
is in operation. The effect of lost motion will be appreciated 
after a study of Fig. 125. Assume that all parts of the 
mechanism are tight except the crank-pin end of the con- 
necting rod as shown. If the engine is rotated by hand, 




Fig. 125. 



for instance, by turning the fly-wheel, the crank will pull 
the piston mechanism and the piston will be drawn into the 
position shown in the upper half of the figure when the crank 
has turned through an angle a. On the other hand, when 
the engine is operating under steam, the piston will push 
the crank pin around and will occupy a position such as 
that shown in the lower half of the figure when the crank 
has been turned through the same angle a. Obviously, 
the piston can occupy two very different positions for the 
same crank position, and a valve setting based upon the 
conditions shown in the upper part of the figure might be 



THE D-SLIDE VALVE 189 

very incorrect when used under the conditions shown in 
the lower part of the figure. 

Lost motion in any part of the mechanism can produce 
analogous results and it is therefore necessary to remove as 
much of it as possible before attempting to set the valve. 
It is practically impossible to eliminate all lost motion, as 
there must be sufficient clearance at all bearing surfaces 
to accommodate a film of oil, and this alone would make 
necessary the taking of indicator diagrams for the check- 
ing of valve settings, even if it were possible to set perfectly 
by measurement for stationary conditions. 

In general, there are two adjustments which can be made 
in setting a plain slide valve. The length of the valve 
stem or eccentric rod can be changed and the eccentric 
can be shifted around the shaft. It is necessary to under- 
stand the effects of each of these adjustments. 

Changing the length of the valve stem is equivalent to 
shifting the valve upon its seat without 
moving the engine as shown in Fig. 
126. In this figure the valve is shown 
in its central position by full lines. The 
lap is the same at both ends. If, now, Fig. 126. 

the valve is worked to the right upon 
its stem by adjustment of the nuts shown, until it reaches 
the dotted position, the head-end lap will have been de- 
creased and the crank-end lap will have been increased by 
the same amount. This would make admission earlier and 
cut-off later for the head end and admission later and cut- 
off earlier for the crank end. Obviously, the effects of 
changing the length of the valve stem are opposite for the 
two ends of the cylinder. 

Shifting the eccentric about the shaft simply changes 
the time relation between valve motion and piston motion; 
it does not alter the valve motion itself. If difficulty is 
experienced in realizing the truth of this statement, it is 
only necessary to draw several Bilgram diagrams for the 




190 STEAM POWER 

same valve, but with different angles of advance, and then 
to construct indicator diagrams for both cylinder ends in 
every case. It will be discovered that shifting the eccentric 
ahead in the direction of rotation, for instance, will make 
all events occur earlier with respect to piston position for 
both ends of the cylinder. 

In setting a plain slide valve which is built symmetrical 
about a central axis, i.e., same inside and outside lap at 
each end, it is first necessary to adjust the length of the valve 
stem. This may be done by removing the steam-chest 
cover so as to expose the valve and then rotating the engine 
slowly by hand and observing the distance traveled by the 
valve on each side of its central position. This is con- 
veniently done by observing the distance between the outer 
edge of the steam port and the outer edge of the valve when 
the valve is fully open at each end. If the valve travels 
further toward the head end than it does toward the crank 
end, with reference to the port edges, the valve stem must 
be shortened; if it travels further toward the crank end 
the stem must be lengthened. 

In making these adjustments it is advisable to turn the 
engine only in the direction in which it is going to rotate, so 
that any lost motion in the valve mechanism will have 
approximately the same effect as when the engine is opera- 
ting. 

When the length of the valve stem is correctly adjusted, 
the eccentric must be so set on the shaft as to give the proper 
angle of advance. This is commonly done by shifting it 
about the shaft until the proper value for the steam lead has 
been obtained. In order to determine the value of the lead 
it is necessary to be able to set the engine on each dead center. 
This can be done approximately by turning the engine until 
the crosshead has come to either end of its stroke, but it 
will be found by trial that the fly-wheel and shaft can be 
turned through a very large angle at each end of the stroke 
without causing an appreciable motion of the crosshead, 



THE D-SLIDE YALYE 



191 



so that this method is not very satisfactory for the purpose 
of adjusting the eccentric. It is customary, therefore, 
to work in such a way as to give a more accurate determina- 
tion of shaft and crank positions for dead center. 

The engine is rotated until the crosshead has been 
brought near one end of its stroke, as shown in Fig. 127, 
and a mark is then scribed across the crosshead and guide 
as at ab. An arc xy is then marked on the fly-wheel by 
means of a tram such as that shown, the end c being placed 




mfrmw 



Fig. 127. 



at point P on some solid part of foundation or floor. The 
engine is then rotated, clockwise in the figure, until the 
crosshead has reached the end of its stroke and returned 
to such a point that the marks on crosshead and guides 
again coincide, as shown by dotted positions in the figure. 
The arc x'y' is then scribed on the fly-wheel with the tram, 
the end c again bearing on the point P. A point z is then 
found by bisecting the arc ef and when this point is brought 
under point d of the tram the crank will obviously be at 
crank-end dead center and the piston at the crank end 



192 



STEAM POWER 




(a) Perfect Cards for Slide Valve Type. 




(b) Actual Card; Small Engine. Center Line of Valve on Center 
Line of Seat; Eccentric Advanced to Give Normal Lead of 
0.05 inch. Engine Running Over. 




(c) Same Setting as (6) except Engine Running Under. 
Fig. 128. 



THE D-SLIDE VALVE 



193 




(d) Angular Advance of Eccentric Increased. Valve Stem Length 
Same as in (6) and (c). Lead 0,375 Inch, 




(e) Angular Advance of Eccentric Decreased so as to Give Negative 
Lead of 0.5 Inch. Length of Valve Stem Unchanged. 




(/) Length of Valve Stem Changed; Angle of Advance as in (6). 

Fig. 128. 



194 STEAM POWER 

of its stroke. A point on the fly-wheel diametrically opposite 
to z is next found, so that when it is brought under point 
d of the tram the engine will be on head-end dead center. 

It is probable that more accurate results are obtained 
by rotating the engine in a direction opposite to that in 
which it rotates under steam, because lost motion is then 
taken up in the same direction as when working, but when 
the whole process of valve-setting is considered it is ques- 
tionable whether this is the correct direction of rotation. 
Opinion and practice differ in this respect. In the end, 
the setting should be checked by the taking of indicator 
diagrams, so that effects of incorrectible lost motion may be 
finally eliminated. 

With the dead -center points found the engine is placed 
on, say, head-end dead center, and the eccentric shifted until 
the valve is open to steam by the desired lead. The eccen- 
tric is then fastened in this position and the engine turned 
to the opposite dead center. Because of angularity of con- 
nections and of irregularities in valve and seat dimensions, 
it generally will be discovered that the valve is not now 
open to steam by the same amount as at the other end. 
If it is desired that it should be, the valve can be shifted 
on its stem about half of the distance by which it is out 
and the eccentric can then be swung about the shaft to take 
up the remaining distance. The effect should then be 
checked by putting the engine on the opposite dead center. 
Valves may be set for equal leads as above, or for equal 
cut-offs or for any sort of a compromise desired. In any 
case the procedure is about the same. The length of the 
valve stem is adjusted, then the eccentric position is 
adjusted, and then refinements are effected by small changes 
of both adjustments. Remember always, that changing 
the length of the valve stem changes events at opposite 
cylinder ends in opposite directions, while shifting the 
eccentric changes all events in the same direction. 

The effects of various adjustments are shown by the 



THE D-SLIDE VALVE 195 

indicator diagrams given in Fig. 128. These diagrams were 
taken from a small, slide-valve engine and serve very well 
to show the way in which the indicator discloses poor 
adjustments. 

PROBLEMS 

1. Given: angle of advance, 30°; throw of eccentric, 1J 
ins.; lead, r§- in.; maximum exhaust-port opening, 1^ in.; find 
the steam lap, maximum opening to live steam, and the exhaust 
lap. 

2. Given: steam lap of £ in.; lead of xe in.; exhaust lap of 
| in.; and the angle of advance equal to 30°. Find the valve 
travel ( = 2 X throw of eccentric) and maximum port opening to 
steam and to exhaust. 

3. An engine has an eccentric throw of If ins.; a steam lap 
of f in.; and a lead of y§ in. Compression begins at J of the 
return stroke. Assume a connecting rod of infinite length and 
find the angle of advance, the exhaust lap, and the maximum 
port openings to steam and to exhaust. 

4. Given: valve travel, 3 ins.; steam lap, f in.; exhaust lap, 
\ in. ; and lead, £ in. ; find maximum port opening, angle of advance, 
and piston positions at cut-off, release, compression, and admission 
for both ends of cylinder, with the length of the connecting rod 
equal to 4J times the length of the crank. 

5. It is required to build an engine having a steam-port opening 
of f in., a lead of -^ in., and a connecting rod four times the length 
of the crank. . Cut-off must occur at f stroke and release at 95% 
of the stroke. Find the inside and outside lap, the throw of the 
eccentric and the fraction of stroke completed by the beginning 
of compression, 



CHAPTER XI 



CORLISS AND OTHER HIGH-EFFICIENCY ENGINES 




Fig. 129. 



90. The Trip-cut-off Corliss Engine. The slide valve 
has certain limitations which can be partly, but never 
wholly, overcome. In most slide-valve gears, for instance, 
the various events occur more slowly than is desirable, 

and this is particularly 
true of cut-off. Ideal 
valves would open sud- 
denly to full opening 
when necessary and 
would close as suddenly 
at the proper time, and 
such action would give 
minimum throttling loss 
and rounding of corners 
of the diagram. Engines 
fitted with such ideal valves would therefore give indicator 
diagrams with maximum work area as shown by the 
dotted lines in Fig. 129, the full lines indicating the type 
of diagram obtained with the ordinary slide valve. 

Again, the simpler forms of slide valve involve the use 
of long ports connecting with the clearance space within 
the cylinder, thus adding greatly to the clearance surface 
exposed and to the cylinder condensation. These ports 
serve for both admission and exhaust, and their walls are 
therefore periodically cooled by the exhaust steam with the 
result that excessive condensation occurs during admission. 
Many attempts have been made to devise valve gears 
which should not be subject to the limitations of the 

196 



HIGH-EFFICIENCY ENGINES 197 

simple slide valve. Some of these have resulted in the 
development of the more complicated slide valves de- 
scribed in the last chapter, but such designs generally leave 
much to be desired. One of the earliest and most success- 
ful solutions was made by Corliss, who developed what is 
known as the trip-cut-off Corliss gear. 

The long combined steam and exhaust ports are elimi- 
nated by the use of four valves, two for steam and two for 
exhaust. These are rocking valves and are located top and 
bottom, at the extreme ends of the cylinder, with their 
longitudinal axes perpendicular to those of the cylinder, 
as shown in Figs. 50, 51 and 52. The exhaust valves are 
located below so as to drain out water of condensation. 
Details of valves of this type are shown in Fig. 130. 

These valves may each be regarded as an elementary 
slide valve which has a cylindrical instead of a flat face, 
and which is oscillated about a center near the face instead 
of being reciprocated, i.e., oscillated about a center at an 
infinite distance. 

The valves are operated as shown in Fig. 131 by short 
links from a wrist-plate pivoted on the side of the cylinder 
and rocked back and forth about its center by means of an 
eccentric operating through the linkage indicated. The 
locations of the various pins and the lengths of the various 
links are so chosen that the valves travel at high velocity 
when opening and closing, that they open very wide, and 
that they close only far enough to prevent leakage and then 
remain practically stationary until about to open again. 
Throttling losses are thus decreased and wear caused by 
useless motion after closure is minimized. 

The opening of the admission valves in this gear is 
effected positively by the linkage already explained, but 
they are closed differently.. For opening, the steam link 
rotates the bell crank B in Fig. 132 and thus raises the 
latch C. The hook on the end of one of the arms of this 
latch engages the steam arm which is fastened on the end 



198 



STEAM POWER 



I 





HIGH-EFFICIENCY ENGINES 



199 




200 



STEAM POWER 



of a rod which is slotted into the end of the valve. The 
valve is thus drawn further open as the wrist plate revolves, 
until the tripping end D of the latch strikes the cam indicated 
by E. This throws the hook out of engagement and thus 
disconnects the valve from the driving mechanism. The 




Fig. 132.— Details of Corliss Trip-Cut-off Gear. 

valve is closed by the action of a dash pot, one form of which 
is shown in Fig. 131. As the steam arm rises during the 
opening of the valve it draws up the plunger or piston of the 
dash pot, leaving a partial vacuum beneath it, and, when the 



D"x 48"Heavy Duty Corliss 
110 Lb. Steam 
R.F.M. 




Fig. 133. 



valve is released by unhooking of the latch, atmospheric 
pressure drives the plunger down and thus causes cut-off 
to occur. The action of a dash pot is found to be unsatis- 
factory when the speed of the engine exceeds about 125 
R.P.M. and most Corliss engines with trip-cut-off operate 



HIGH-EFFICIENCY ENGINES 201 

at still lower speeds. Under such circumstances the cut- 
off is very rapid as compared with the piston speed, and the 
diagram shows a comparatively sharp corner at this point. 
A set of diagrams obtained from a large Corliss engine 
operating at 80 R.P.M. is given in Fig. 133, and it is obvious 
that little throttling occurs. 

Because of the low speed at which these engines operate 
the stroke can be made long with respect to the diameter 
without attaining a prohibitive piston speed. The economy 
mentioned in Chapter VII as resulting from the use of long 
strokes can thus be obtained in these engines. An idea 
of the saving in steam effected by the partial elimination 
of throttling and condensation losses by means of the 
Corliss gear can be obtained from the curves in Fig. 134 
(a) and (6), which give average performances. 

The position of the cam which determines the time 
at which cut-off occurs is controlled by the governor of 
the engine. When moved in the direction taken by the 
steam arm it causes cut-off to occur later. Variation of 
the point of cut-off is used in these and in most other engines 
to control the amount of work done per cycle in order that 
the engine may make available the quantity demanded at 
the shaft, as will be explained in a later chapter. It is there- 
fore desirable that the range of cut-off should be as great 
as possible, but it has been found very difficult to design 
trip-cut-off gears which will give a cut-off later than about 
0.4 stroke if steam and exhaust valves are operated from the 
same eccentric. Later cut-off causes poor timing of the 
exhaust events. 

This has led to the introduction of Corliss engines 
with two eccentrics and two wrist plates per cylinder. 
One set operates the steam valves and the other the exhaust 
valves. With this arrangement the range of cut-off is 
unlimited. 

91. Non-detaching Corliss Gears. Because of the low 
speed at which trip-cut-off Corliss engines are operated,, 



202 



STEAM POWER 













, 
























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03 XT1 ,-y 

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3 PR0XIMATE STEAM CONSUMPTIO 

OF 

VARIOUS TYPES OF ENGINES 

(NON-CONDENSING) 

nple Single-Valve Throttling Engines (P=100 1 
nple Single-Valve Automatic Engines (P = 100 I 
nple Four- Valve Automatic Engines (P=100 Lb 
ndem and Cross-Compound Four- Valve and 

Corliss Engines(P=100 lbs.) 
ndem and Cross-Compound Four- Valve and 

Corliss Engines(P=125 lbs.) 
ndem and Cross-Compound Four- Valve and 

Corliss Engines(P=150 lbs.) 
itz Engine P=133 lbs.; 92.7° Superheat 
.5H.P.; R.P.M. 206 
aflow Engine, P =150 lbs. 
aflow Engine, P = 140 lbs.; Superheat, 110°F. 








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STEAM POWER 



they are necessarily large, heavy and costly and efforts 
have been made to design gears which shall possess the 
advantages of the original Corliss mechanism without 
the limitation as to speed. 

In many models the Corliss valves are retained and are 
located in the ends of the cylinder as just described or in 




Fig. 135. — Non-detaching Corliss Valves Located in Cylinder Head. 



the cylinder heads as shown in Fig. 135. In some the wrist 
plate and the connecting links are also retained, but in 
others they are eliminated. In all engines of this type the 
admission valves are closed positively, the closure being 
effected by the same linkage that opens the valves to admit 
steam. Quick action is obtained by the arrangement of 
the operating mechanisms, the centers of rotation and the 



HIGH-EFFICIENCY ENGINES 205 

lengths of links being so chosen that the valve travel is 
small when the valves are closed, that it is rapid when the 
valves are opening and closing, and that the valves remain 
practically wide open during most of the time that steam 
is being admitted. 

The advantages of small clearance and short and sepa- 
rate ports are attained in these arrangements and the 
operation of the valves is almost as perfect as that of the 
trip-cut-off gear. Engines fitted with these modified 
Corliss gears are operated at speeds considerably higher 
than those permissible with the older arrangement, and they 
may be classed with medium-speed engines. 

Engines of this type are generally known commercially 
as four- valve engines, but as this name applies equally well 
to the ordinary trip-cut-off gear and to others which will 
be described later, it is best to use some other designation. 
The term non-detaching Corliss engines seems to best 
describe them and is apparently gaining in favor. 

Non-detaching Corliss engines generally give diagrams 
intermediate between those obtained with the low-speed, 
trip-cut-off mechanism and those obtained from slide-valve 
engines with the simpler forms of valves, though the later 
designs very closely approximate the performances of the 
trip-cut-off Corliss engine. 

92. Poppet Valves. Attention has already been called 
to the fact that the use of highly superheated steam is 
very effective in lessening or even eliminating initial con- 
densation. Experience has shown that large valves and 
valves with sliding surfaces such as slide valves and Corliss 
valves do not work well with highly superheated steam. 
The large castings warp so that contact surfaces do not 
remain true and the lack of moisture which acts as a seal 
with saturated steam leads to excessive leakage. Dif- 
ficulty has also been experienced with the lubrication of 
these sliding types of valves when using highly superheated 
steam. 



STEAM POWER 




An old form of 
valve known as the 
poppet valve has re- 
cently been adopted 
by some builders as 
a solution of the 
difficulties met in 
the use of highly 
superheated steam. 
This form of valve in 
four-valve arrange- 
ment, combined with 
designs in which short 
ports and symmet- 
rical cylinder cast- 
ings are used, yields 
very economical en- 
gines which can be 
safely used with a 
degree of superheat 
prohibitively high in 
the case of the slid- 
ing and oscillating 
forms of valves. 




Fig, 136&. — Cross-section, Lentz 
Engine. 



HIGH-EFFICIENCY ENGINES 



207 



Sections of a modern type of poppet valve engine are 
shown in Figs. 136 (a) and 136 (6), and details of the 
admission valve and its operating mechanism are given in 
Fig. 137 (a) and (b). The valves are all double-seated 
(double-ported or double-beat), that is, they seat at both 
ends and are made hollow so that the steam passes both 
around the outside of the valve and through the valve 
as shown by the arrows in Fig. 137 (b). This results in 
large area for passage of steam and in quick opening and 



Rolle; 
Cam Surfa 




To Cylinder 



Fig. 137a. — Admission Valve and Operating 
Mechanism, Lentz Engine. 



Fig. 1376. 



closing, as in the case of gridiron valves, with small actual 
movement of the valve. 

The valves are opened positively by eccentrics opera- 
ting through cams and rollers as shown in Fig. 136 (6) and 
they are closed by springs as rapidly as the return motion 
of the cam permits. The eccentrics are mounted on a 
horizontal lay shaft which is located to one side of the 
engine, with its axis parallel to that of the latter, and which 
is driven by bevel gears from the crank shaft of the engine. 

Since this valve arrangement gives short steam and 
exhaust poits, permits the use of small clearance, and 



208 STEAM POWER 

gives fairly rapid opening and closing of valves with little 
throttling when open, it gives good economy when used 
with saturated steam. By adding superheat the economy 
is still further improved. The water rate of one of these 
engines is shown for one load in Fig. 134 (a). A simple, 
Lentz non-condensing engine is reported to have given 
a consumption of 16.13 lbs. of steam per horse-power hour 
with 92.7° superheat, and a pressure of 133 lbs., and this 
figure is materially lowered by compounding, higher super- 
heat, lower back pressure, etc. 

93. The Una-flow Engine. A very interesting modifica- 
tion of the steam engine, known as the Una-flow Engine, 
has recently appeared. In this design an attempt is made 
to decrease the loss due to condenastion in a very original 
way and the results of tests seem to indicate that the design 
makes possible very great economy. 

In the ordinary forms of engine the entire wall of the 
cylinder is subjected to the cooling action of the lowest 
temperature steam during the entire exhaust stroke, and in 
double-acting types these cooled walls are immediately 
brought into contact with the higher pressure steam acting 
on the other side of the piston, as well as coming into con- 
tact later with the next charge of high-pressure steam. 
The uDa-flow design minimizes this action by admitting 
steam at the ends of the cylinder, exhausting it at the 
center of length of the cylinder, and compressing the steam 
caught in the clearance up to a value approximating initial 
pressure, thus heating the clearance walls. The heating 
of the clearance walls is further effected by partly jacket- 
ing the head with live steam on its way to the admission 
valve and the jacket is sometimes extended along the 
cylinder to the point at which cut-off normally occurs. 

One form of this engine is shown in Figs. 138 and 139. 
The steam enters the cylinder head from below, passes 
up to a double-seated poppet valve, flows into the cylinder 
until cut-off occurs and then expands until the piston 






HIGH-EFFICIENCY ENGINES 



209 



uncovers the exhaust ports. The steam is exhausted 
until the returning piston again covers these ports, after 
which the material trapped within the cylinder is compressed 
as indicated in the diagrams. The ideal sought is to main- 
tain each part of the wall approximately at the tempera- 
ture which the expanding steam will have when reaching 
it and thus to minimize thermal interchanges and loss. 




Fig. 138. — Section of Una-flow Engine Cylinder. 



Tests of these engines show that very great ratios 
of expansion can be used in a single cylinder without the 
excessive losses customary when such ratios are attempted 
in the ordinary counter-flow type. It is thus possible to 
obtain good economy with one una-flow cylinder expanding 
from a high pressure to a vacuum; conditions which would 
involve the use of compounding with ordinary construc- 
tion. 



210 



STEAM POWER 



The results of tests on one of the first Una-flow engines 
built in this country are shown in Fig. 134 (a) and (b). In 
comparing with the curves it should be noted that two of 
the tests were run with high superheat. 

94. The Locomobile Type. In the effort to improve 
the ecomony of small steam plants the Germans developed 
a form of plant now known as the Locomobile Type. The 




Fig. 139. 



name came from the fact that these plants, as originally 
made, were mounted on wheels and intended for portable 
use by agriculturists and contractors. Their economy in 
the use of fuel proved so great that they have since been 
built for stationary use in sizes running well toward 1000 
horse-power per unit. 

A locomobile of American construction known as the 
Buckeye-mobile is illustrated in Fig. 140, which shows a 
longitudinal section of the plant. The tandem compound 






HIGH-EFFICIENCY ENGINES 



211 




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212- STEAM TOWER 

engine is mounted on top of an internally fired boiler with 
the engine cylinders located in the flues which lead the 
products of combustion away from the boiler. 

The steam generated in the boiler is passed through 
a superheater suspended in the smoke box. The flow of 
steam is from the rear toward the front of this superheater 
(counter flow) so that the hottest steam comes in contact 
with the hottest gas. The steam then passes through a 
pipe contained within the flue to the high-pressure cylinder, 
which is jacketed by the hot flue gases and in which the 
loss of heat to metal is thus minimized. From the high- 
pressure cylinder the steam passes to a receiver contained 
in the smoke box, the receiver serving as a reheater to 
evaporate any condensate exhausted from the first cylinder 
and to superheat the steam admitted to the low-pressure 
cylinder. From the low-pressure cylinder, the steam 
passes through a feed-water heater in which it raises the tem- 
perature of the boiler feed and then it passes to atmosphere 
or to a condenser. Boiler-feed pump and condenser pump, 
if used, are also integral parts of the plant, being driven 
directly from the main engine. 

It will be observed that every precaution is taken to 
guard against initial condensation, and to minimize loss 
of heat in flue gases and in exhaust steam leaving the 
plant. The high economies achieved are due to such 
facts alone. 

Small plants of this type have given an indicated horse- 
power hour on a little over one pound of coal when oper- 
ated condensing, whereas the best large compound recip- 
rocating engine plants seldom do better than about 1.75 
lbs. of coal per I.h.p. and often use 2 or more pounds when 
operated condensing, 



CHAPTER XII 
REGULATION 

95. Kinds of Regulation. There are two distinctively 
different kinds of regulation referred to in connection 
with reciprocating steam engines, one of which may be 
called fly-wheel-regulation and the other governor-regula- 
tion or governing. 

The regulating effect of the fly-wheel has already been 
referred to. The turning effort exerted at the crank pin 
by the action of steam on the piston or pistons of an engine 
is not constant, and the angvlar velocity of the engine shaft 
is therefore constantly varying during each revolution. 
It is the function of the fly-wheel to damp these variations 
so that they do not exceed the allowable maximum for 
any given set of operating conditions. The efficiency of 
the fly-wheel in this respect is measured by the coefficient 
of fly-wheel regulation b w which is defined by the equation 

IT T7 

, (66) 



y 

in which 

T 7 max = maximum velocity attained by a point on 

fly-wheel rim or other revolving part; 
Fmm = minimum velocity of the same point, and 
F = mean velocity of the same point 

' max I * min • , i 

= —jr— - approximately. 

Governor-regulation is absolutely different. Its function 
is to proportion the power made available to the instan- 
taneous demand. The fly-wheel takes care of variations 

213 



214 



STE 



VOW^U 



occurring during the progress of one cycle, while governor 
regulation varies the work value of successive cycles. 

96. Governor Regulation. If the effect of engine 
friction be neglected, the power delivered at the shaft of 
the engine will vary directly with the indicated horse- 
power. Such an assumption is accurate enough for the 
discussion which follows. 

The indicated horse-power of a given engine is deter- 
mined entirely by the value of the mean effective pressure 
and the number of cycles produced in a given time, since 
these are the only variables in the formula for indicated 
horse-power. The power made available by an engine 
might therefore be varied by varying the mean effective 





Fig. 141.— Throttling 
Governing. 



Fig. 142.— Cut-off 
Governing. 



pressure, or by varying the number of cycles produced in 
a given time, or by a combination of both processes. 

All of these possibilities are used. In ordinary station- 
ary power plants the mean effective pressure is generally 
varied. In the case of pumping engines, working against 
a constant head, but required to deliver different quantities 
of water at different times, the number of cycles per minute 
is generally altered by changing the speed at which the 
engine operates. In locomotive and hoisting practice 
both the number of cycles per minute (speed) and the 
mean effective pressure are varied as required to meet 
the instantaneous demands. 

These variations may be effected manually as by the 
driver of a locomotive, in which case the engine may be 
said to be manually governed. Or, they may be brought 



REGULATION 215 

about mechanically, as in the case of most stationary power- 
plant engines, in which case the engine may be said to be 
mechanically governed. In some instances a combination 
of manual and mechanical governing is used. 

97. Methods of Varying Mean Effective Pressure. The 
mean effective pressure increases and decreases with the 
area of an indicator diagram of constant length, so that 
the mean effective pressure can be changed by any method 
which will change the area of the diagram. Two methods 
are in use and they are illustrated in Figs. 141 and 142. 
The first causes a variation in area by changing the value 
of the initial pressure. This is generally done by chang- 
ing the opening of a valve in the steam line just outside of 
the steam chest. It is called throttling governing, and the 
valve is called a throttling or throttle valve. The latter 
name is also commonly used for the valve located near the 
engine, which is used to shut off the supply of steam entirely 
when the engine is not in operation. 

The second method, illustrated in Fig. 142, is known 
as cut-off governing. The variation of cut-off determines 
the amount of steam admitted to the cylinder per cycle 
and is used to measure out the quantity required for the load 
which happens to exist at any instant. Cut-off governing 
is used on most modern stationary engines and is exclusively 
used in large reciprocating engine power plants. 

98. Constant Speed Governing. Most engines used for 
such purposes as the operation of mills and the driving of 
electrical and centrifugal machinery are required to run 
at practically constant speed irrespective of the load. They 
are furnished with mechanical governors which so regulate 
the power made available that there shall never be any 
appreciable excess or deficiency which would respectively 
cause an increase or a decrease in speed. 

These mechanical devices always contain some sort 
of tachometer which moves whenever the speed of the engine 
exceeds or falls below the proper value. The tachometer 



216 STEAM POWER 

is so connected to the valve gear that it decreases the 
power-making ability of the engine whenever the speed 
starts to increase and it increases the power-making ability 
if the speed drops. 

Since the valve gear must have a different position 
for each load in order that it may throttle or cut off as 
necessary to suit that load, it follows that the tachometer 
which controls the position of the valve gear must also 
have different positions for different loads. But tachom- 
eters assume positions dependent on speed, and therefore 
different loads can only be obtained if the tachometer and 
the engine to which it is connected operate at different 
speeds for different loads. 

Constant-speed governing is therefore an anomaly. 
The device which is supposed to maintain constant speed 
irrespective of load must be operated at different speeds, 
as the load varies, in order that it may maintain the valve 
gear in the different positions required to handle the differ- 
ent loads. All so-called constant-speed engines have their 
highest speed when carrying no load, and the speed gradually 
decreases to a minimum as the load increases to a maxi- 
mum. The total variation is generally between 2 and 4%. 

The efficiency of a governor in this respect is measured 
by means of the coefficient of governor regulation, 8 G , 
which is defined by the equation 

&g= , (67) 

in which 

ri2 = highest rotative speed attained by the engine; 

rt\ = lowest rotative speed attained by the engine, and 

n = mean speed 

n 2 +ni . . 

= — - — approximately. 

99. Governors. The mechanical devices which are 
used for controlling the power-making ability of an engine 



REGULATION 



217 



as described above are known as governors. There are 
many varieties and only a few of the more prominent can 
be described. 

(a) The Pendulum Governor. One of the earliest forms 
of governor used on steam engines is illustrated in Fig. 
143. It is often called a 
fly-ball governor. This 
governor is driven by 
gearing, chain or belt 
from the engine, and the 
weights assume some 
definite position for 
each different speed, thus 
drawing the collar to 
different positions. The 
valve gear is connected 
to this collar and is 
moved correspondingly. 

A similar governor 
is shown in Fig. 131, yig. 143. 

which also indicates the 

way in which the collar is connected to the valve gear in 
the Corliss type of engine. The governor rods are moved 
as the collar moves and they in turn alter the position 
of the knock-off cam, and thus vary the time at which 
cut-off occurs. As the speed increases due to a decrease 
of load, the governor weights and collar move up, and this 
shifts the cams so as to produce earlier cut-off and decrease 
power-making ability. 

(b) Shaft Governors. On medium- and high-speed 
engines fitted with some form of slide valve it is found best 
to use what are known as shaft governors. They are gen- 
erally carried within the fly-wheel of the engine, operate in 
a plane passing through the rim of the wheel at right angles 
to the shaft, and operate upon the eccentric in such a way 
as to vary the cut-off with speed (and load) changes. 




218 



STEAM POWER 




Fig. 144. 



One simple form of such a governor is shown in Fig. 
144. The eccentric is not mounted directly upon the 

engine shaft, but is carried 
by a pin P in the fly-wheel 
and is slotted so that it can 
swing back and forth across 
the shaft, about P as a center. 
Its position at any time is 
determined by the position 
of the governor weights W, 
which draw the eccentric 
down (in the figure) as they 
move out. 

The center of the eccentric 
is indicated by a heavy dot 
in the figure, and it will be seen that this center would 
travel in the arc of a circle about P, as the weights moved. 
If the path of the eccentric center is drawn on a Bilgram 
diagram, it will be found that this motion is equivalent to 
decreasing the length of the eccentric crank and increasing 
the angle of advance, resulting in earlier cut-off as the 
weights move out with increasing speed and decreasing 
load. Other events will also be changed as the eccentric 
swings, and some of these changes are occasionally unde- 
sirable. 

Numerous designs have been developed in which the 
eccentric is so guided as to produce various sorts of rela- 
tions between the different steam and exhaust events. 
All can be divided into two classes, those in which the 
eccentric swings about a fixed center variously located, and 
those in which the center of the eccentric is guided to 
move in a straight line. All can be studied by plotting 
the path of the eccentric center (path of Q) on the Bilgram 
diagram. 

The Rites Inertia Governor is a form of shaft governor 
so designed as to act very quickly with change of speed, 



REGULATION 



219 



and to be very powerful, so that it can shift heavy parts. 
It is shown in place in the wheel in Fig. 145. With changes in 
speed it acts like a governor of the type just described, 
swinging (with increasing speed) about a fixed point P in 
the wheel as its center of gravity G moves outward under 
the action of the centrifugal effect C and against the 
action of the spring. This motion shifts the center of the 




Fig. 145. 



eccentric from E toward e, giving the desired variation in 
cut-off. 

Superposed upon this action is that of inertia. Assume 
the wheel and governor to be rotating clockwise at a given 
constant speed. If the engine speed is suddenly increased, 
the wheel will move faster, but the governor bar will tend 
to continue rotating at the same speed because of its inertia. 
It will thus lag behind the wheel, rotating about P and bring- 
ing about an earlier cut-off. The position thus assumed 
will later be maintained by centrifugal effect if the new speed 



220 STEAM POWER 

is maintained. The particular advantage resulting from 
using inertia in this way is speed of action. In many forms 
of governor the inertia of the moving parts actually resists 
the efforts of the governor to assume the new position 
required by chauged load and speed. 



CHAPTER XIII 

THE STEAM TURBINE 

100. The Impulse Turbine. One of the oldest of modern 
water wheels is the tangential or impulse wheel shown 
diagrammatically in Fig. 146. Water flowing from a 
reservoir above the wheel passes through a nozzle and the 




Fig. 146. — Tangential or Impulse Wheel. 

jet, moving at high velocity, strikes buckets on the rim 
of the wheel and causes the latter to revolve. Theoretically 
the velocity of the water in the jet would be 

v = V2gh feet per second, .... (68) 
in which 

g = gravitational constant, 32.2, and 
h = head in feet as shown in the figure. 

The kinetic energy possessed by the moving water would 
be 



w v 



221 



(69) 



222 



STEAM POWER 



in which w represents pounds of water discharged per second 
and g and v have the same meanings as above. 

If the buckets of the wheel could reduce the velocity 
of the water to zero they would absorb all of this kinetic 
energy and (assuming no losses within the buckets and the 
bearings of the wheel) would make all of it available at 
the shaft for the doing of useful work. 

Any fluid moving at velocity v and striking buckets in 
the form of a jet would possess kinetic energy in quantity 
given by Eq. (69) and would drive the wheel in the same 
way. Steam might therefore be used instead of water 
with exactly the same results, and steam is so used in what 
are known as impulse steam turbines. 

Experience shows that steam will flow at high velocity 

from any opening made in the 
steam space of a boiler or 
from any open-ended pipe con- 
nected to such a boiler. This 
is commonly said to be due 
to the high pressure within 
the boiler, the spectator pic- 
turing the process as the 
driving out of part of the 
steam by the high-pressure steam within the boiler, just as 
though the part leaving were a solid piston and were driven 
out as is the piston of an engine during 
admission, as shown in Fig. 147. 

An hydraulic analogy is given in Fig. 
148. The vessel shown is supposed to be 
fitted with a piston, and it is assumed to be possible to exert 
any desired pressure upon the piston. Any such pressure 
exerted is the exact equivalent of some given head of water 
and the resultant jet velocity would be given by Eq. (68) 
by substituting for h the head in feet equivalent to the 
pressure exerted upon the piston. 

When an " elastic " fluid such as steam is being con- 




Fig. 147. 



Fig. 148. 



THE STEAM TURBINE 223 

sidered it is, however, necessary to take account of other 
factors. The steam within the boiler exists at a high 
pressure; after issuing it exists in the atmosphere at a lower 
pressure. But low-pressure steam contains less heat than 
does steam at high pressure, and this difference must exist 
in some form, as it is energy and could not possibly have been 
destroyed during the flow. 

Experiment shows that steam after flowing into the 
atmosphere from a boiler in this way has exactly the same 
characteristics as though it had expanded adiabatically 
behind a piston through the same temperature range, ex- 
cepting for the fact that it has a very high velocity, which 
it would not possess if expanded behind a piston. Experi- 
ment further shows that, if small losses be neglected, the 
kinetic energy possessed by a jet of steam is exactly equal 
to the energy which would be turned into work if that steam 
acted on a piston as in an ordinary engine. 

A complete picture of the process of flow can then 
be made by assuming the steam flowing out in the form of 
a piston driven by high-pressure steam, as before, and adding 
to this the idea that this piston expands adiabatically as 
it travels from the region of high to that of low pressure. 
This expansion liberates heat contained within the piston 
or plug of steam and this heat is used in imparting addi- 
tional velocity to the moving steam which is giving up this 
heat. 

The result of using such a jet upon a theoretically 
perfect tangential or impulse wheel would be to rob the 
jet of all this energy. But the energy possessed per pound 
of steam in the jet is just the same as that shown under the 
upper lines of a complete expansion cycle using one pound 
of steam. The area under the upper horizontal line of the 
PF-diagram of the cycle as shown in Fig. 21 may be assumed 
to represent the work done upon one pound of steam (flow- 
ing out) by another pound which is being evaporated and 
pushing out the first in order to make room for itself. The 



224 



STEAM POWER 



area under the expansion curve in the PF-diagram repre- 
sents the energy converted into velocity energy by the adia- 
batic expansion of the flowing steam. The lower hori- 
zontal line represents the negative work during condensa- 
tion to water at the lowest pressure and temperature, and 
the left-hand line represents the pumping of this water back 
into the boiler and the raising of its temperature to the value 




Fig. 149. — Early Form of Impulse Turbine. 



maintained within the boiler. The complete expansion cycle 
is therefore the cycle upon which the impulse steam turbine 
operates and, as a matter of fact, it is the theoretical cycle 
of all steam turbines. 

The ideal impulse turbine would therefore be acted 
upon by a jet which possessed available kinetic energy 
represented by the area of the complete expansion cycle. 
If the buckets could entirely remove this energy, that is, 
could reduce the velocity of the jet to zero, the same amount 



THE STEAM TURBINE 



225 



of energy could theoretically be made available at the shaft 
of the turbine. 

An example of a simple form of impulse steam turbine 
is given in Fig. 149, in which the essential parts of an early 
form of Kerr turbine are shown. The wheel, the diaphragm 
and nozzles are all inclosed within a casing. The space 
on one side of the diaphragm is connected to the steam pipe 
and that on the other is in communication with the space 
into which the exhaust steam is to be exhausted. 

Another form of impulse turbine is shown in Fig. 157. 
It will be described later. 

101. Theoretical Cycle of Steam Turbine. It was shown 
in the preceding section that the steam turbine operates 
on the complete expansion cycle. If a turbine could remove 
from the steam passing through it and convert into mechan- 
ical form all of the energy which is theoretically possible, 
it would therefore make available mechanical energy 
represented by the area of the PF-diagram of the complete 
expansion cycle. The area of the corresponding T<t>- 
iiagram would show the 
game quantity measured in 
thermal units. The theory 
of .the steam turbine can 
therefore be studied by 
means of these two dia- 
grams. 

In Fig. 150 is shown the 
T(£-diagram of the complete 
expansion cycle for several 
different conditions. The 
figure abed represents con- 
ditions when the steam is 
dry and saturated at the beginning of the adiabatic ex- 
pansion cd. Constant quality lines are designated by x 
and x'. It is obvious that by the time the steam has 
expanded down to the pressure at d it will have a quality 




Fig. 150. 



226 STEAM POWER 

less than unity. If, therefore, it be in the form of a jet 
issuing from a nozzle and having a high velocity by virtue 
of its acliabatic expansion, the jet will really be a mixture 
of steam and water. 

If the steam be superheated at constant pressure as shown 
by ce before passing through the nozzle, it is evident from the 
figure that the jet issuing from the nozzle will contain 
less water than in the preceding case, because the condition 
of the material in the jet after adiabatic expansion will be 
as shown at / instead of as shown at d. The cycle in such 
a case would also be larger by an amount indicated by the 
area cefd, representing just that much more heat converted 
into mechanical energy per pound of steam or other unit 
for which the diagram happened to be drawn. 

If superheating had been carried to the point indicated 
by g before expansion, the jet would obviously issue from 
a nozzle in the form of superheated steam as shown by the 
point h in the figure. In that case the cycle would be 
abcgha, and superheat would have to be removed from the 
low-pressure steam to bring it to the conditions indicated 
at i before condensation could begin. 

If desired, the PF-diagrams for such cycles can be drawn 
very easily. The line be, or be or bg is a horizontal line in 
the PF-diagram. The line ha is similarly horizontal and the 
line ab is vertical. The adiabatic expansion is represented 
by a curved line in the PF-diagrams, but can be drawn 
easily because the necessary data are obtainable from the 
T^-diagram, in which this expansion is represented by a 
straight line. 

ILLUSTRATIVE PROBLEM 

Draw the PF-diagram for a steam turbine receiving one pound 
of steam at a pressure of 200 lbs. absolute, with a tempera- 
ture of 500° F. and exhausting against a pressure of 0.5 lbs. 
absolute. 

First, locate on a TV-chart for steam the point representing 
the condition of steam at 200 lbs. pressure with a temperature 



THE STEAM TURBINE 



227 



of 500° F., and draw a vertical line extending downward 
until it cuts the horizontal temperature line corresponding to 





































73 


























































































| 


















1 

• 
/ 




































/ 


















/ 
















/ 


















/ 




J 


i 

— 




__-•— " 




I—" 






8 



§*3 

B 
U 

i 2 



>o 



•sqyuibS J3d 'sqi-oanssoj^ 



0.5 lb. pressure. This is practically at 540° F. absolute, or about 
80° F. 



228 STEAM POWER 

Second, take from the steam table the volumes of one pound 
of steam at, say, 200 lbs., 140 lbs., and 100 lbs. absolute pressure 
when superheated to the values shown by this vertical line. These 
will be about 2.75 cu.ft., 3.58 cu.ft., and 4.67 cu.ft., respectively. 
Plot these volumes with corresponding pressures on a PF-chart 
as shown in Fig. 151. 

Third, take from the TV-chart the pressures at which the 
vertical line intersects different volume lines in the wet steam 
region and plot volumes against pressures on the TV-chart. 

Fourth, draw a smooth curve, as shown, through all points 
so determined. 

Fifth, draw horizontal top and bottom lines and a vertical 
line at the left of the diagram. This vertical line should be to 
the right of the pressure axis by an amount representing the 
volume of" one pound of water, but the volume is so small that 
it cannot be plotted to any ordinary scale. 

102. Nozzle Design. It was stated in preceding sec- 
tions that the energy which would be converted into work 
by the introduction and adiabatic expansion of steam 
behind a piston is converted into kinetic energy when steam 
flows out of an orifice or nozzle and that an ideal impulse 
turbine could absorb all this kinetic energy from the jet, 
bringing it to rest and making the energy available in the 
form of useful power at its shaft. It is, therefore, of interest 
to determine the velocity which a jet will acquire under 
different conditions. 

This could be done by evaluating the area of a diagram, 
such as that of Fig. 151, and then putting this value ivi 
place of K in Eq. (69) and solving for v, but it can be done 
much more accurately and expeditiously in other ways. 
The heat energy which can be converted into kinetic 
energy of the moving jet and which can later be con- 
verted into useful work by the turbine wheel is represented 
by the area enclosed within the lines of the complete 
expansion cycle when drawn on the T^-diagrams. That 
is the area abed in Fig. 152, for instance, for the case of 
wet steam at the beginning of expansion. But this area 
is equal to that representing the heat supplied minus 



THE STEAM TUKI3INK 



229 



that representing the heat rejected, that is, Q1-Q2, so 
that 

K(inB.t.u.)=Qi-Q2 (70) 

The values of Qi and Q 2 can be found very readily by 
plotting the points c and d upon a 7>-chart for steam and 
observing the constant heat lines upon which they fall, 



It 


b 

1 , 


c 


\ 


/ 


f 






\ 


..> 




h / 






h 








N 


1 \ 





Fig. 151 



or they can be obtained even more conveniently from what 
is known as a Mollier Chart for steam. In this chart, 
entropy above 32° F. is plotted against heat above 32° F. 
as shown in Fig. 153. An adiabatic expansion on this chart 
is shown by a horizontal line, since this shows a constant 
entropy change just as a vertical line on the 7> chart shows 
a constant entropy change. 

If a point is found in this chart giving conditions corre* 
sponding to those at point c in Fig. 152, the value of Q x 



230 



STEAM POWER 




THE STEAM TURBINE 231 

can be read directly under that point on the horizontal 
axis. A horizontal line drawn from that point to the 
terminal-pressure line will give the point corresponding to 
d of Fig. 152 and the value of Q2 can be read on the hori- 
zontal axis immediately below that point. The difference 
between the two readings gives the value of the kinetic 
energy K or of the mechanical energy which an ideal tur- 
bine could make available, but the expression will be in 
British thermal units and not in foot-pounds. 

This value of the kinetic energy, i.e., K = Qi — Q2, may 
then be placed in Eq. (69), giving, 

K = 77S(Q 1 -Q 2 )=~itAbs. ) . . . (71) 

since Qi and Q2 refer to one pound of steam, or 

K = 77Sw(Q-Q 2 )=^ L ftAb8., . . . (72; 

when w represents the number of pounds of steam flowing 
per second. 

Solving cither Eq. (71) or Eq. (72) for v gives, 

v = Vf7SX2g(Q 1 -Q 2 ) 

= V778X64.4(Q!-Q 2 ) 

= 22WQ X -Q2 feet per second. . . (73) 

The design of a nozzle consists simply in choosing such 
sections that the desired amount of steam may flow through 
it with the desired pressure drop, as the velocity obviously 
is determined by that pressure drop. This is very con- 
veniently done by working in terms of one pound of steam, 
since all formulas and charts are generally given on that 
basis, and then multiplying the cross-sectional areas found 
by the number of pounds of steam required. 

Assume for instance that it is desired to design a nozzle 



232 STEAM POWER 

to pass one pound of steam per second with an initial pres- 
sure of 100 lbs. per square inch abs., and a terminal 
pressure of 60 lbs., the steam being initially dry and satu- 
rated. 

The Mollier chart shows that Qi is equal to about 1187 
B.t.u. per pound of steam, while Qo is equal to about 1147 
B.t.u. The velocity with which a jet would issue from a 
theoretically perfect nozzle under these conditions may 
then be found by using Eq. (73). This gives 

v = 22Wl 187 -1147 
= 1416 feet per second. 

The shape of the entrance end of the nozzle is generally 
made such that the steam will enter it without great dis- 
turbance and the shape beyond that point is determined 
by methods which will be explained below. The cross- 
section of the discharge end must be such as to pass the 
required quantity at the velocity found above to be equal 
to 1416 feet per second. This is easily done by deter- 
mining the volume of steam discharged. 

Drawing the adiabatic expansion on the T$-chart will 
give the quality at the end of the expansion; or, the quality 
can be determined by finding what quality a pound of 
steam at 60 lbs. pressure must have to give it a heat content 
of 1147 as found above. With the quality known the ter- 
minal volume per pound can be found by multiplying the 
quality by the specific volume at terminal conditions. 
Thus for the case under discussion the quality will be 
about 96.7% and as the specific volume at 60 lbs. is 7.17 
cubic feet, the volume to be passed per second, per pound 
of steam is 0.967X7.17 = 6.94 cu.ft approximately. If the 
velocity is 1416 feet per second the area per pound of steam 
must be 6.94-^1416 = 0.0049 sq.ft. 

The exact shape of the nozzle is determined by deciding 
upon the way in which pressure, or velocity, or volume 



THE STEAM TURBINE 



233 



shall change as the steam passes through it. Suppose, for 
instance, that a nozzle is to be constructed of the length 
shown by ab in Fig. 154, and that the pressure is to vary 
along its length as shown. Assume also that the nozzle 
is to pass 10 lbs. of steam per second. Taking initial 
pressure as 100 lbs. and terminal as 60 lbs., the conditions 



a 

^100 
c 
w 
* 90 



so 



I 60 

























■■»» 














a- 
















b 








-- -- ■ 




























", 


i 




, 


'2 


























\> 


h 


















% 


> y 


















Xi^ 


fes 


**b. 










/ 








<^fj 










/ 


















/va 


riati 


dii o: 


Velocity 










1 




















T 


en art 


hof 


Nozz 


e 







1400 
1200 
1000 
800 
600 
400 
200 



Fig. 154. — Nozzle Design. 

will be the same as in the problem above. The discharge 
area will have to be 10X0.0049 sq.ft. or 0.049 sq.ft. 

The area at the plane X2 must be that required to pass 
the steam when it has the velocity resulting from expansion 
from 100 down to 64 lbs., just as though the nozzle ended at 
that point. This can be found just as the terminal area 
was found above. Similarly the sections at X\ and x can 
be found by figuring velocity and area for expansions to 



234 



STEAM POWER 



74 and 90 lbs., respectively. If the various areas required 
are determined in this way, the nozzle will have a longi- 
tudinal section about as shown by the dotted lines in the 
figure and the variation of velocity will be about as shown 
by the curve. 

If the shape of a nozzle is determined in the same way 
for a case in which the terminal pressure is less than about 
0.58 of the initial pressure, the nozzle will be found to have 

a very different shape. This is 
shown in Fig. 155. The nozzle 
is known as an expanding nozzle 
and the smallest section is known 
as the neck. The pressure P„ 
in the neck is always equal to 
about 0.58 Pi and the velocity 
in the neck is always equal to 
just over 1400 feet per second. 
It is therefore the section at 
the neck which determines the 
quantity of steam which a nozzle 
will discharge if expanding to a 
pressure equal to or lower than 
0.58 Pi. 

103. Action of Steam on 

Fig. 155.-Expanding Nozzle. ^^ ^^ R hag ^ 

stated that the steam acting in an impulse type of turbine 
delivers energy to the wheel of the turbine by giving up 
its kinetic energy. In an ideal turbine the steam jet would 
be brought to rest and would thus give up all of its kinetic 
energy. 

In real turbines it is impossible to bring the jet to rest, 
as practical design problems prevent it. There is there- 
fore always a loss in real machines because of the residual 
or terminal velocity of the steam as it leaves the wheel. 
Thus let the black section in Fig. 156 represent the section 
of a bucket or blade sticking out radially from the rim of a 




THE STEAM TURBINE 



235 



wheel, the wheel revolving about the axis indicated by the 

dot dash line but located behind the plane of the paper. 

If minimum loss by eddying is to be experienced at the 

point at which the steam jet 

enters the blade, the jet must 

enter the blade along a tangent 

to the curve of the inside of the 

blade at the entrance edge. This 

direction is shown by the line 

marked v r in the figure. 

Were the bucket stationary, 
the steam jet would move as 
shown byr r , but as the bucket 
moves ahead, and, so to speak, 
runs away from the jet, the 
steam must really travel in a 
direction such as that indicated 
by v a in order to strike the 
bucket in the direction indicated 
by v T . The conditions governing 
the flow of steam into a bucket 
are the same as those governing the speed with which and 
direction in which an individual runs toward and jumps 
upon a moving vehicle. He will experience least shock 
when he is moving ahead at the same rate as the vehicle 
at the instant when he gets on board. His motion must 
therefore be made up of two, one toward the vehicle and 
the other in the direction of the vehicle's travel. 

In the case of steam flowing onto a blade as shown in 
Fig. 156, the various velocities are so related that when 
drawn to scale the real or absolute velocity of the steam, 
v a , and the real or absolute velocity of the blade, v b , form 
two sides of a triangle of which the closing side represents 
v r , the velocity of the steam relative to the bucket. The 
value and direction of v r is obviously obtained from v a by 
geometrically subtracting the velocity of the bucket. 




Fig. 156 



236 STEAM POWER 

After entrance, the steam flows around the inner curve 
of the blade and is finally discharged with the same rela- 
tive velocity as that with which it entered, and at an angle 
set by the tangent to the inner curvature of the discharge 
edge of the blade as shown by v R . But, since the steam 
has been moving ahead with the same velocity as the 
bucket during the entire time that it was in contact with 
the bucket, it is also moving ahead with a velocity v b when 
it leaves the wheel. Its real or absolute velocity is then 
v A , which is found by combining vr and v b as shown in the 
figure. 

The kinetic energy possessed by the jet when entering 

the blade is equal to ~~- ft.-lbs., and that which it possesses 
when leaving is — j= — •. Obviously, the energy removed 

01)1) " 11)1) A 

while passing over the blade is ~ - — •. If the blade 

2g 2g 

were theoretically perfect, it would be so constructed that 
v A 2 would be zero and all of the kinetic energy would then 
be removed. This is practically impossible in a real mechan- 
ism, and there is always a loss due to the residual velocity 
v A . The best that can be done is to so choose the angle 
of jet and blade, and the velocity of blade with respect 
to the steam that the actual numerical value of v A is made 
as small as possible. 

Designs usually work out in such a way that this occurs 
when the blade velocity is equal to about 0.47 of the abso" 
lute velocity of the steam jet. 

104. De Laval Impulse Turbine. The expanding nozzle 
already described was first used by De Laval in an impulse 
type of turbine. The essential elements of this device are 
shown in Fig. 157. The nozzles are arranged at such an 
angle to the plane of the wheel that the steam jets strike 
radially arranged blades at the proper angle to enter without 
much loss. The blades deflect the jets as shown and 



THE STEAM TURBINE 



23; 



absorb the greater part of their kinetic energy, so that 
the steam flows away from the wheel with low absolute 
velocity. 

As many nozzles are used as are required to make avail- 



Nozzle 



Steam !Tn 




iTufbine Shaft 



^ ^jNo'zzles- 
Fig. 157. — Single Stage, De Laval Impulse Turbine. 

able the amount of energy desired at full load, and pro- 
vision is made for shutting off one or more nozzles by hand 
when conditions do not warrant the use of all. Governing 
for ordinary variations of load is effected by throttling 
the steam flowing to the nozzles in use, thus altering the 
initial pressure as necessary. 



238 STEAM POWER 

A section through the wheel and casing of such a tur- 
bine directly connected to a centrifugal pump is given 
in Fig. 158. The steam flows into the live steam space 
through a throttle valve controlled by the governor; the 
valve and connections are not shown in the illustration. 
From the live steam space the steam flows through nozzles 
not shown, and into the exhaust steam space, thus acquir- 
ing a high velocity. The buckets of the wheel are located 
just in front of the discharge ends of the nozzles and the 
steam moving at high velocity must pass through them 
before moving on toward the exhaust outlet. 

105. Gearing and Staging. It has been stated that the 
most efficient operation with ordinary designs is obtained 
when the blade speed is equal to about 0.47 of the absolute 
steam velocity or, roughly, half the velocity of the imping- 
ing jet. To get high economy in the use of steam, large 
pressure drops are used and very high jet velocities result. 
When the buckets of a turbine are operated at peripheral 
speeds equal to half these jet velocities one of two diffi- 
culties is often met. The stresses induced in the wheel 
structure by centrifugal effects become so high as to offer 
serious difficulties in design, or the rotative speed of the 
unit becomes too high for direct connection to the machine 
which is to be driven. 

One method of partly overcoming the latter difficulty 
is to operate the turbine at or near the theoretically desir- 
able speed and transmit the power to the driven machine 
through gears which decrease the rotative speed to the 
necessary extent. This method was used with all of the 
early De Laval turbines which were of comparatively small 
capacity. It is now being successfully applied to marine 
propulsion and other purposes for which large units are 
used. It is only a partial remedy in the case of large units, 
however, as the gears necessary for the desired reduction 
and the size of the turbine wheels would both become 
excessive. 



THE STEAM TURBINE 



239 




a> 

-a 

c 



240 STEAM POWER 

Another and very common method is known as com- 
pounding or staging. This may be of two varieties. The 
pressure drop in each stage may be limited to that 
which will give a reasonable velocity and a number of 
such stages may be put together in series on one shaft. 
This would give one set of nozzles and a wheel for each 
stage, the steam discharged from one wheel with very 
low velocity expanding to a lower pressure through the 
nozzles of the next stage and impinging upon the 
wheel of that stage with the resultant high velocity. 
Such an arrangement is known as pressure staging or 
pressure compounding, and is extensively used in large 
turbines. 

The pressure staging method is illustrated in Fig. 159 
as applied to the De Laval type of impulse turbine. The 
combined increase in diameter of wheels and increase 
in length of blades gives the necessary increase in area to 
pass the larger volumes of steam as the drop of pressure 
continues from stage to stage. 

Instead of staging on a pressure basis, staging on a veloc- 
ity basis may be used. In such a case the drop in pressure 
through one set of nozzles is great and the resultant veloc- 
ity high. The steam moving at this high velocity is then 
directed upon the buckets moving at such peripheral velocity 
that they absorb only part of the kinetic energy of the steam, 
discharging it with a lower absolute velocity than that 
with which it entered, but one which is too high to be 
thrown away. The steam then passes through a set of 
stationary vanes which direct it upon the blades of a second 
wheel, in passing through which it gives up still more of 
its kinetic energy with a corresponding further decrease 
of velocity. If the velocity still possessed by the steam 
warrants it, a second set of stationary guide vanes and a 
third set of moving buckets can be supplied for further 
reducing it and by carrying this velocity staging through 
a sufficiently great number of stages any initial velocity 



THE STEAM TURBINE 



241 



could be absorbed theoretically without the use of wheels 
with high peripheral speeds. Practically, losses due to 




friction, eddying and other sources limit the number of 
velocity stages to two or three. 



242 



STEAM POWER 




Fig. 160.— Early Form of Curtis Turbine. 



THE STEAM TURBINE 



243 



Velocity staging is combined with pressure staging 
in the Curtis type of turbine. A section through part 
of an early design of vertical turbine of this t} r pe is shown 
in Fig. 160. The turbine illustrated had four pressure stages 
and each pressure stage had two velocity stages. 

Many varieties of impulse turbines have been developed 
and all of the larger ones employ several wheels and sets 
of nozzles and diaphragms to obtain the necessary staging. 
The same result has been obtained in some of the smaller 
models by discharging the steam from nozzles on to a set 
of buckets which are able to absorb only a fraction of the 
kinetic energy, catching it at discharge and returning it 
for another passage through the buckets, and so on until 
the greatest practical fraction of the kinetic energy has been 
absorbed . 

106o The Reaction Type. If high-pressure steam or 
other fluid be forced into a de- 
vice arranged as shown in Fig. 
161 and free to revolve about 
a vertical axis, the jets blowing out 
of the nozzles will cause the mecha- 
nism to revolve in the direction 
indicated by the arrow. This rota- 
tion is said to be due to the reaction 
of the jets, and the mechanism there- 
fore constitutes a simple form of reaction turbine. By 
increasing the number of nozzles 
any amount of steam could be dis- 
charged and therefore any amount 
of work could be obtained. 

This multiplication of nozzles 
can, however, be more conveniently 
accomplished by fastening radial 
vanes to the periphery of a wheel 
as shown in Fig. 162, the space 
between any two vanes constituting a nozzle through which 




Fig. 161. 

Elementary Reaction 

Turbine. 




Fig. 162. 



244 STEAM POWER 

the steam can discharge. By mounting such a wheel 

within a casing as shown in Fig. 163 it forms a simple 

reaction turbine. One of the characteristic differences 

^- between the impulse and the reaction 

types lies in the distribution of pressures. 



In the impulse type the nozzles are 

fastened into a stationary part of the 

HE<3 turbine and the drop of pressure occurs 



b 






k * entirely within the nozzles. The wheels 
Sa/v are therefore immersed in a space in 
rmz^J which a uniform, low pressure exists. 
^ In the reaction type, on the other hand, 
the nozzles are carried on the wheel and 
there must be a higher pressure on one side of the wheel 
than there is on the other. Since there must also be me- 
chanical clearance between the blade tips and the interior 
of the casing, it follows that the reaction type will be 
handicapped by considerable leakage which does not exist 
in the impulse type, excepting as some of the jet may 
" spill " over the ends of the blades. 

The difference of pressure on the two sides of the wheel 
also causes a tendency toward motion of the wheel along 
the shaft, or of the wheel and shaft, in a direction away from 
the higher pressure. 

Many unsuccessful efforts have been made to design 
efficient reaction turbines, but no pure reaction type has 
yet been commercialized. The turbines commonly called 
reaction turbines are really combinations of reaction and 
impulse types. 

One example of what is commercially called a reaction 
turbine is shown in Fig. 164. Alternate rings (or rows) 
of stationary and movable blades guide the steam as it 
expands from the high pressure at one end to the low pres- 
sure at the other. The stationary blades project inward 
from the interior surface of the stationary casing and the 
movable blades project outward from the external surface 



THE STEAM TURBINE 



245 




246 



STEAM POWER 



of the cylindrical rotor. The rotor blades act like those 
of an impulse turbine in partly reversing the direction of jets 
of steam which reach them with comparatively high veloci- 
ties, but they also act like the movable nozzles el a reac- 
tion turbine since the steam in passing through them expands 
and acquires kinetic energy, giving a reaction on discharge. 
The stationary blades serve to redirect the steam so that 
it strikes the next set of moving blades at the proper angle 

and they also serve as 
nozzles in which velocity 
energy is acquired. This 
is shown diagrammatic ally 
in Fig. 165, in which S 
denotes stationary, and M 
movable blades. 

The Parsons type, il- 
lustrated in Fig. 164, may 
be described as a multistage 
type in which impulse and 
reaction are" utilized in con- 
junction. 

The balance pistons 
shown in the figure are 
used to balance the end 
thrust caused by the differ- 
ence in pressure existing on 
opposite sides of the wheels 
in the case of reaction turbines. Each piston is of such 
a diameter that it presents a surface equal to the blade 
surface acted upon by one of the unbalanced pressures, 
and by connecting across as shown in the figure a high 
degree of balance is secured. 

The overload valve is used to admit high-pressure 
steam to the low-pressure blades for carrying excessive 
overloads. The larger area of the passages through these 
blades permits an abnormal amount of high-pressure steam 




THE STEAM TURBINE 247 

to pass, thus giving a high load-cany ing capacity with 
decreased economy. 

107. Combined Types. The clearance at the ends of 
the stationary and moving blades in the Parsons type of 
turbine permits considerable steam to leak by, as previously 
explained. This clearance must have almost the same 
length (measured from blade tip to opposing metal) in all 
stages in order to insure freedom from rubbing, but it is 
more detrimental in the high-pressure stages than in the 
low. The high-pressure blades are much shorter than the 
low-pressure blades and a leakage length of a certain amount 
is therefore equal to a greater fraction of the total blade 
length. The density of the high-pressure steam is also so 
much greater than that of the low-pressure steam that many 
more pounds can leak through an opening of a given size 
in a given time. In discussions of this character, it should 
not be forgotten, however, that leakage area is determined 
by the dimension already referred to multiplied into a 
circumference and that the circumference is much greater 
at the lower end. 

Because of these and other reasons many manufacturers 
have come to the conclusion that the impulse type is best 
for the high-pressure end of the turbine and the reaction 
type for the low-pressure end. Many such combinations 
have been produced and they are giving very good results. 

108. Economy of Steam Turbines. In general, the 
economies of steam turbines and reciprocating engines are 
about the same when each type is operated at normal load 
and under the best conditions. It is probable that very 
large turbines have a slight advantage over reciprocating 
engines (as generally built) in the matter of economy and 
the reverse of this statement appears to be true for most 
small units, although very economical turbines have been 
produced in small sizes in the past few years. 

The turbine, however, generally gives a flatter water- 
rate curve than does a reciprocating engine; that is, for 



248 STEAM POWER 

loads each side of the most economical the steam per horse- 
power hour does not increase above the value attained 
at most economical load as rapidly in the case of turbines 
as it does in the case of most reciprocating engines. With 
a very variable load, therefore, or with a load which is far 
removed from the rated value, the turbine probably gives 
a better average performance than does the reciprocating 
engine. This is particularly true in large sizes. 

It has been shown that the turbine operates on the complete- 
expansion cycle and it will be remembered that the recipro- 
cating engine operates on a cycle with incomplete expansion. 
The turbine is therefore able to make better use of very 
low-pressure steam than can the piston type. 

Trial on a T<f> or Mollier chart will show that a turbine 
receiving steam at about atmospheric pressure and expand- 
ing it down to a vacuum of from 28 to 29 ins. should make 
available as much work as one receiving steam at a high 
boiler pressure and expanding down to atmospheric. In 
other words a drop of 100 lbs. or more above atmospheric 
pressure makes no more energy available than does a drop 
of about 13 lbs. below atmospheric, or the lower the initial 
pressure the more heat is converted into work by a given 
pressure drop. A small decrease in back pressure (terminal 
or condenser pressure) is therefore very effective in the case 
of turbines. Tests show that an increase of one inch of 
vacuum will cause an increase of economy of from 3 to 10 
per cent, depending upon the type of turbine and upon 
other factors. 

Experience has shown that reciprocating engines are 
fully the equal of turbines in the high-pressure ranges, 
in many cases they are even superior, but the turbine is far 
superior in the low-pressure region and in cases where very 
great ratios of expansion are to be used. Advantage has 
been taken of the superior ability of the turbine to handle 
low-pressure steam by constructing mixed plants, recipro- 
cating engines being used for expanding down to the neigh- 



THE STEAM TURBINE 249 

borhood of atmospheric pressure and turbines expanding 
the steam exhausted by these engines to the lowest vacuum 
which can be maintained economically. This system has 
been found particularly useful for increasing the capacity 
of a reciprocating-engine plant. The capacity of such a 
plant can often be almost doubled without any increase 
in boiler capacity by simply inserting turbines into the 
exhaust lines between the engines and the condensers, and 
then arranging the pressures so that the turbines carry 
the expansion from about 16 lbs. absolute down to a vacuum 
of from 28 to 29 ins. Turbines used in this way are called 
low-pressure or exhaust-steam turbines. 

Superheat is also very effective in bettering turbine 
economy, every ten degrees of superheat generally causing 
a saving of about 1 per cent in the weight of steam required 
per horse-power. 

The steam turbine is generally cheaper than the recipro- 
cating engine of like capacity if the conditions of operation 
permit the use of the high rotative speed characteristic 
of the turbine. It is therefore extensively used for direct 
connection to blowers, centrifugal pumps and electrical 
machinery. Most of the larger electric power stations 
which have been installed within the past few years have 
used turbines to drive the generators, and single units 
direct connected to 20,000 K.W. generators are now numer- 
ous. Units rated at 45,000 K.W. are also being installed 
and units of still larger capacity have been designed. 

PROBLEMS 

1. A steam turbine produces one horse-power hour at its shaft 
for every 30 lbs. of steam supplied. The initial pressure is 200 
lbs. absolute and the steam is superheated 200° F. The turbine 
exhausts against a back pressure of 14 lbs. absolute. 

Find the thermal efficiency on the assumption that heat 
of liquid at exhaust temperature is not chargeable to the turbine. 

2. Develop a complete expansion cycle for one pound of 
material used under the conditions of Prob. 1 and find the energy 



250 STEAM POWER 

made available per cycle. From this value determine the number 
of pounds of material theoretically required per horse-power hour 
and compare with the value given in Prob. 1. 

3. Find the additional quantity of energy which would theoret- 
ically be made available per pound of steam in above problems if 
the back pressure could be lowered to \ lb. absolute. 

4. Develop a complete expansion cycle from an initial pressure 
of 225 lbs. absolute with a superheat of 200° F. to a back pressure 
of \ lb. absolute. Assume that this is to be divided up into six 
parts, each making available the same quantity of energy. Find 
the pressure drop for each part. Note that this is most easily 
done with the help of the Mollier chart. 

5. A steam turbine receives steam at a pressure of 225 lbs, 
per square inch absolute and with a superheat of 190° F. and 
exhausts into a condenser in which a pressure of J lb. per square 
inch absolute is maintained. The turbine is direct connected to 
an electric generator and produces a K.W.-hour on 12 lbs. of steam. 
If a K.W.-hour is equivalent to 3411 B.t.u., what is the thermal 
efficiency of the combination? 

6. Develop a complete expansion cycle for the conditions of 
Prob. 5 and determine the pounds of steam which would be re- 
quired theoretically to develop energy equivalent to 1 K.W.-hour. 
Compare with the value given in Prob. 5. 

7. Determine the velocity theoretically attainable by expanding 
steam in one step from the initial to the final conditions of Prob. 
5 above. What would be the value of the kinetic energy of such 
a jet per pound of steam flowing? 

8. Determine the shape of a nozzle required to discharge 
1000 lbs. of steam per hour, initial conditions being 100 lbs. per 
square inch absolute, and dry saturated steam; final pressure 
being 2 lbs. absolute. 

9. Determine velocity and kinetic energy of jet in Prob. 8. 



CHAPTER XIV 
CONDENSERS AND RELATED APPARATUS 

109. The Advantage of Condensing. The lowest pres- 
sure which can be used in a steam-engine cylinder, that is 
the exhaust pressure, is determined by the pressure prevail- 
ing in the space into which the steam is exhausted. With a 
given initial pressure the amount of work which can be ob- 




Fig. 166. 

tained theoretically from a given weight of steam increases 
as the exhaust or back pressure decreases, as shown by the 
areas of the two diagrams in Fig. 166, and experience has 
shown that at least a part of this theoretical increase can be 
obtained in real engines. It is therefore desirable to ex- 
haust into a space in which the lowest possible pressure 
exists when the work obtained per pound of steam is the 
only consideration. 

The most available space into which an engine can 

251 



252 STEAM POWER 

exhaust is that surrounding the earth and already occupied 
by the earth's atmosphere. The pressure in this space 
is approximately equal to 14.7 lbs. per square inch at sea 
level and is due to the weight of the atmosphere. Since 
the superincumbent column of atmosphere decreases in 
depth as one moves upward, its weight also decreases and 
atmospheric pressure therefore averages less than 14.7 
lbs. per square inch at high altitudes and has a greater 
average value at points below sea level. 

If it is desired to exhaust into a pressure lower than 
atmospheric a means of maintaining such an abnormal 
pressure within some sort of vessel must be devised. It 
is the purpose of a condenser and its associated apparatus 
to make available a space in which such a low pressure can 
be maintained. Its method of operation will be considered 
in later sections. 

There is also another advantage which may be ob- 
tained by the use of a condenser. It often happens that 
the water available is not well adapted to use in boilers. 
It may be salt water as in marine practice, or it may contain 
a number of undesirable gases and solids in solution as often 
occurs in stationary practice. Some types of condensing 
apparatus are so arranged that the steam exhausted from the 
engine is converted into liquid water without admixture 
and can therefore be returned to the boiler as practically 
pure water, thus largely eliminating the troubles that 
would ensue from the use of poor feed water. 

110. Measurement of Vacuums. Assume that some 
non-volatile liquid, that is, a liquid that did not vaporize, 
could be found and also that it contained no gases in solu- 
tion. If a long tube were inserted in a vessel filled with 
such a liquid and had its upper end connected with some 
form of vacuum pump which could remove air from its 
interior, as shown in Fig. 167, liquid would rise in the tube 
as the air was removed. Removal of air would result in 
lowering the pressure within the tube, but the constant 



CONDENSERS AND RELATED APPARATUS 253 



C^\ 



Ik 



-Pa 



atmospheric pressure on the liquid surface outside the tube 

would then force liquid up the tube to such a height that 

the pressure p a of air in the tube 

plus the pressure due to the column 

of liquid of height h within the 

tube just equaled the pressure due 

to the atmosphere on the surface 

of the liquid in the vessel. If the 

pump could remove all of the air 

from the tube, liquid would rise to 

such a height that the pressure 

exerted by it on a plane passing 

through the lower surface just 

equaled that of the external atmos- 
phere. 

The same result could be at- 
tained by using a tube closed at 

one end, filling it with the liquid, 

and then inverting so that the end 

rested' in the liquid as shown in 

Fig. 168. If the tube were long 

n enough, the liquid would drop to some such 

point as shown, under which conditions the 
height of liquid would just balance atmos- 
pheric pressure. This would only be true if 
the liquid did not volatilize and did not contain 
gases in solution; with these assumptions the 
space above the liquid in the tube would con- 
tain absolutely nothing. This space would be 
said to be perfectly vacuous, or a perfect vacuum 
would be said to exist in that part of the tube. 
A device of this character is used to 
measure the pressure of the atmosphere and 
is known as a barometer. Mercury is used 

as the liquid because its high density makes it possible 

to use a short tube and because it may be considered 



Fig. 167. 



Fig. 168. 



254 



STEAM POWER 



as non-volatile at ordinary temperatures. The average 
atmospheric pressure at sea level, equal to 14.7 lbs. per 
square inch approximately, can support about 30 ins. 
of mercury, so that this figure is generally taken as the 
standard sea level barometer reading. An atmospheric 
pressure of less than 14.7 lbs. would be shown by a barom- 
eter reading of less than 30 ins.; a greater atmospheric 
pressure by more than 30 ins. Corresponding values of 
atmospheric pressure and barometer reading are given 
in Table VII. To this have also been added the altitudes 
to which the different values would correspond if a pressure 
of 14.7 lbs. existed at sea level and there were no variations 
of atmospheric pressure excepting those due to change of 
elevation. Values of this type can only be roughly approx- 
imate, because local barometric variations are constantly 
occurring and the sea-level atmospheric pressure varies 
both sides of 14.7 lbs. 

TABLE VII 
Atmospheric Pressure, Barometer Reading and Altitude 

(Negative signs mean distance below sea level.) 



Barometer, 


Atmospheric Pressure, 


Altitude, 


Inches of Mercury. 


Pounds per Square Inch. 


Feet (Approximate). 


25.00 


12.27 


4750 


26.00 


12.76 




26.50 


13.01 


3250 


27.00 


13.25 




27.50 


13.49 


2250 


28.00 


13.74 




28.50 


13.98 


1300 


29.00 


14.23 




29.25 


14.35 




29.50 


14.47 


450 


29.75 


14.60 




30.00 


14.72 


Sea level 


30.25 


14.84 




30.50 


14.96 


-450 


30.75 


15.09 




31.00 


15.21 


-900 



CONDENSERS AND RELATED APPARATUS 255 

The exact value of standard atmospheric pressure 
at sea level is taken at 29.921 ins. of mercury, which is 
equal to 14.696 lbs. per square inch and corresponds to 
the 760 mm. of mercury, used by scientists as standard. 

A tube with both ends open and arranged as shown 
in Fig. 167 can be used to measure the degree of vacuum 
existing in the space to which its upper end is connected, 
and many vacuum gauges are constructed on this principle, 
using mercury as the liquid. The extent to which the 
pressure is lowered in the top of the tube is indicated by 
the height to which the mercury column rises and this 
height in inches is used as a measure of the vacuum. Thus 
if a perfect vacuum were created and if the atmospheric 
pressure were equal to 14.7 lbs. the gauge would show 
about 30 ins. of mercury above the level in the reservoir. 
If the vacuum were less perfect the gauge would show a 
shorter column. 

It should be noted that the reading of the vacuum 
gauge does not give the pressure existing in the vacuous 
space, but gives the amount by which the pressure has been 
reduced below that of the atmosphere, the difference 
being expressed in inches of mercury. By subtracting this 
reading from the existing atmospheric pressure expressed 
in the same units, the absolute pressure in the partially 
vacuous space (expressed in inches of mercury) is obtained. 

It is obvious, therefore, that a vacuum-gauge reading 
of say 28 ins. of mercury does not always mean the same 
absolute pressure. With a barometer reading of 28 ins. 
it would represent a perfect vacuum; with a barometer 
reading of 30 ins. it would represent a partial vacuum, the 
absolute pressure in the partially vacuous space being 
equal to 2 ins. of mercury. 

111. Conversion of Readings from Inches of Mercury 
to Pounds per Square Inch. It is often necessary to con- 
vert readings of pressure in inches of mercury into pounds 
per square inch. This can be done with sufficient accuracy 



256 STEAM POWER 

under ordinary circumstances by multiplying the inches 
of mercury by the constant 0.4908. Thus, 

Barometer in inches X 0.4908 = atmospheric 

pressure in pounds per square inch . . . (74) 
and 

(Barometer in inches — vacuum gauge in inches) 
X 0.4908 = absolute pressure in partially 
vacuous space in pounds per square inch. . (75) 

ILLUSTRATIVE PROBLEM 

A vacuum gauge constructed like that shown in Fig. 167 
reads 27 ins. when the barometer reads 29.5 ins. What is the 
absolute pressure in the partial vacuum above the mercury? 

The absolute pressure is equal to 29.5—27=2.5 ins. of mer- 
cury, which is equal to 

2.5X0.4908 = 1.227 lbs. per square inch. 

112. Principle of the Condenser. A perfect vacuum 
could be created in any closed vessel with impenetrable 
walls if a pump could be devised which could remove all 
material contained within that vessel. Or, any degree of 
vacuum can be maintained in any partially closed vessel 
by fitting it to a pump which can remove all material 
flowing into the vessel as fast or faster than it enters, raise 
the pressure of this material to atmospheric or higher and 
discharge it. 

The latter principle is made use of in real condensers, 
a pump of some form, or an equivalent, removing from the 
condenser the material exhausted by the engine and in- 
leakage from the atmosphere, and discharging it at atmos- 
pheric pressure at a sufficiently rapid rate to maintain 
the desired vacuum. If the condenser and connections 
can be made leak proof, the pump or equivalent has to handle 
only the material exhausted from the engine. 

A steam engine exhausts a mixture of steam, water 
and gases, the gases being a mixture of those originally 



CONDENSERS AND RELATED APPARATUS 257 

dissolved in the boiler-feed water and air which leaks into 
those parts of the system in which a partial vacuum is main- 
tained. If the pump had to handle the same volume of 
material as is exhausted by the engine, no gain of work 
would result from condensing, because the pump would 
have to do at least as much work in raising the pressure 
of this material to atmospheric and discharging it as could 
be obtained by allowing it to expand in the engine. 

Steam, however, occupies a much larger volume than 
water at the same temperature and pressure. Thus steam 
at 212° F. occupies a volume of about 26.79 cu.ft. per 
pound, but water at the same temperature and pressure 
occupies a volume of only about 0.0167 cu.ft. per pound; 
at a temperature of 120° F. which is often used in condensers, 
the specific volume of steam is about 203 and that of water 
only 0.0162. Therefore, if the steam is condensed after 
exhaust from the engine and before entering the pump to 
be discharged to atmosphere, the pump work is greatly 
reduced. The volume of the condensate is almost negli- 
gible in comparison with the volume of steam exhausted, 
and the work of pumping it is almost negligible in compari- 
son with the work it made available in the engine. 

Gases contained in the exhaust steam cannot be lique- 
fied and must be pumped as gases. The work required 
to pump them can, however, be reduced by lowering their 
temperature as far as possible. 

The condenser equipment may be regarded as con- 
sisting of a combination of a partially closed vessel and 
some form of pump. The vessel is so constructed that a 
low temperature can be maintained within it and that 
large quantities of heat can be removed from it for the 
purpose of condensing the exhaust steam and of cooling the 
contained gases. This is generally done by using large 
quantities of cool water. 

The absolute pressure within the condenser is made 
up of two parts. The two parts are, (a) that due to the 



258 STEAM POWER 

water vapor, since the space over the condensed water will 
always be filled with saturated steam at the same tempera- 
ture (approximately) as that of the water, and (6) that due 
to any gases present. 

The pressure of the saturated steam (water vapor) 
can be found from the steam tables opposite the temperature 
existing in the condenser and it is the pressure which would 
exist in the condenser of an ideal system in which no gases 
were mixed with the working substance. The pressure 
of the gases can be found by subtracting from the total 
measured pressure in the condenser the pressure exerted 
by the water vapor as shown in the steam tables. The 
pressures exerted by the water vapor and gases are spoken 
of as partial pressures, since their sum makes up the total 
pressure within the condenser. 

The presence of gases causes a two-fold loss. First, 
it increases the pressure against which the engine has to 
exhaust, thus raising the back-pressure line on the diagram 
and decreasing the work area. Second, it increases the 
work which must be done by the pump which otherwise 
would only pump the condensate and such saturated water 
vapor as accompanied it. 

113. Types of Condensers. The condensers actually 
used in steam plants can be roughly divided into two types, 
as 

(a) Contact condensers and 

(6) Non-contact condensers. 
In the first type the water which is used for condensing 
and cooling is intimately mixed with the exhaust from the 
engine within the condensing vessel, and the resultant 
mixture of condensing water, condensate and gases is drawn 
out of this vessel and discharged to atmosphere by the 
pump. 

In the second type condensing water flows on one side 
of metal surfaces of some sort and the exhaust is led over 
the other side, the heat being transmitted through the 



CONDENSERS AND RELATED APPARATUS 259 



Condensing 
Water In. 



from Engine 



metal. In condensers of this type the condensate and 
gases are not mixed with the condensing water and the 
condensate can therefore be returned to the boiler as feed 
water with the advantages already mentioned. 

114. The Jet Condenser. One of the earliest forms of 
contact condensers which is still very widely used for 
moderate vacuums is commonly 
known as the jet condenser. The 
principle of operation of the jet 
condenser is shown in Fig. 169. 
Water, under pressure, entering 
as indicated, is broken up into 
fine streams or jets and sprayed 
into the exhaust coming from 
the engine. The resultant mix- 
ture flows downward into the 
neck of the condensing vessel or 
" condenser head " and is re- 
moved by some form of pump. 
This pump handles gases, vapors 
and water and is known as 
a vacuum pump, a wet-vacuum 
pump, or a ivet-air pump, the 
term wet signifying that it han- 
dles the water as well as the 
gases. 

The pressure within such a 
condenser head would be theo- 
retically equal to that corresponding to the temperature 
of the resultant mixture if no gases were present. In 
practice the pressure of the water vapor would roughly 
correspond to the average temperature near the top of the 
vessel and there would be a partial pressure due to gas 
as well. This gas would consist of that brought over by 
the engine exhaust plus that released from the condensing 
water under the low pressure within the condenser. 




^Mixture to Pump 

Fig. 169. — Jet Condenser. 



260 



STEAM POWEE 



Details of a complete jet condenser and of the method 
of connecting it to an engine are given in Fig. 170. The 
atmospheric relief valve is installed in all condensing sys- 
tems and is arranged to open automatically to atmosphere 
if the pressure within the system rises to atmospheric, that 
is, if the " vacuum is lost." 



* (a) 




Steam 
Cylinder 



Fig. 170. — Jet Condenser and Method of Connecting to Engine. 



With the jet condenser the pressure might start to 
rise because of slow action or even stoppage of the pump. 
As the condenser head filled up the rising water would 
ultimately entirely cover the jet and condensation would 
then practically cease. In the arrangement shown in 
Fig. 170 there is an additional safety device which breaks 
the vacuum in the exhaust system if the water in the head 



CONDENSERS AND RELATED APPARATUS 261 

rises above a certain height, thus preventing the external 
atmospheric pressure from forcing this water back along 
the exhaust pipe and into the cylinder, an event which 
would probably result in a wrecked engine. 

The jet condenser here described is known as a parallel- 
flow type, because everything within the condensing vessel 
flows in the same direction. The gases and vapors handled 
by the pump theoretically have the same temperature 
as that of the mixture with which they flow out at the 
bottom of the condenser head. The temperature of this 
mixture therefore determines the temperature of the gases 
and vapors pumped. 

There are numerous forms of contact condensers which 
more or less closely resemble the types of jet condenser 
just described. They are occasionally all classed as jet 
condensers, but more often are given distinguishing names. 

One very common form of contact condenser is generally 
known as a barometric condenser. It consists essentially 
of a condenser head, similar to that used with the jet con- 
denser already described, and a tail pipe or barometric 
pipe which partly or wholly takes the place of the wet- 
vacuum pump by removing part or all of the mixture formed 
within the condenser. One model of such a condenser is 
shown in Figs. 171 and 172. 

The exhaust from the engine enters the head through 
the large pipe shown and divides into two parts, one part 
passing down through the center of the head and the re- 
mainder flowing downward in the annular space A. The 
condensing or injection water enters as shown and is divided 
by the spraying cone and injected into the engine exhaust, 
which enters the central tube of the condenser. The 
mixture thus formed flows downward and finally meets the 
discharge from the lower end of the annular space A, which 
is then condensed. The mixture of injection and con- 
densing water together with such gases as have been en- 
trapped, then flows downward into the tail pipe, which is 






262 



STEAM POWER 



over 34 ft. in length and which dips into the " hot well " 
at its lower end. As atmospheric pressure can only sup- 



/To Vacuum or 
y Air Pump 



Cooler 



Exhaust 




Wafer for Cooling 
"Air" 



-Drain Injection 

J Water 



Fig. 171. — Barometric Condenser. 



port a column of water about 34 ft. high, the tail pipe forms 
an automatic wet-vacuum pump, water flowing from it as 
rapidly as it accumulates within it. 



CONDENSEES AND RELATED APPARATUS 263 




so 

a 



m 



264 



STEAM POWER 



Atmospheric Relief Valve 



Experience has shown that the maintenance of a high 
vacuum with this type of condenser depends upon the ex- 
tent to which gases are removed from the condenser head. 
These gases are generally called air, as the greater part of 
them is air. In the type illustrated such " air " as is 
not trapped by the descending mixture rises through the 

hollow spraying cone, then 
through the air cooler and 
flows out through the pipe 
indicated to the vacuum or 
dry-air pump. The air in 
rising through the center 
of the spraying cone is 
cooled by the water flowing 
around it, and it is further 
cooled by coming into con- 
tact with water as it works 
its way through the air 
cooler. This results not 
only in lowering its tem- 
perature, but also in caus- 
ing the condensation of a 
great deal of the water 
vapor accompanying it. 
This condensed vapor col- 
lects in "the space surround- 
ing the air cooler and 
flows down into the head 
through the drain shown. The vacuum pump, therefore, 
handles cool gases containing little water vapor and prac- 
tically no liquid water. It is sometimes called a dry-air 
pump or dry- vacuum pump for this reason. 

The entrainer shown in the exhaust system in Fig. 172 
is so shaped that water collecting in the exhaust piping 
and flowing into the entrainer is picked up by the exhaust 
steam and carried into the condenser. 




Fig. 173. 



-Baragwanath Barometric 
Condenser. 



CONDENSERS AND RELATED APPARATUS 265 

The flow of steam and injection water in this condenser 
is parallel, but the material on its way to the dry-vacuum 
pump flows upward and the cooling water flows downward 
so that counter-current flow is used in this part of the appa- 



SYPHONING ITS 
WATER FROM 
£ TANK OR FLUME 



BARAGWANATH 

CONDENSER 

ORDINARY 

SETTING 




Engine 



-^^^e 




Fig. 174. 



ratus. This has the advantage of bringing the air leaving 
the condenser into contact with the cooling water just as 
it enters and therefore when it has its lowest temperature. 

A somewhat similar condenser, arranged so that it 
requires no pump, is shown in Figs. 173 and 174 (a) and (6). 



266 



STEAM POWER 



Exhaust and injection water mix as shown, the quantity 
of injection water being regulated by the hand wheel on 
top of the condenser. The mixture flows downward through 
the narrow neck and the velocity attained in this part of the 
tail pipe is so high that all air and similar gases are swept 
along with the current. 

For starting, the 

"Air" ^' 

valve V in Fig. 174 (b) 
is opened, allowing water 
to flow into the lower 
part of the tail pipe. 
This creates a partial 
vacuum, and atmos- 
pheric pressure then 
forces water up the in~ 
jection pipe and into 
the condenser head. The 
valve V is then closed 
and the condenser con- 
tinues to siphon its own 
water. Because of this 
action this type is often 
called a siphon conden- 
ser. By supplying a 
circulating pump as in- 
dicated in Fig. 174 (a) 
it can be converted into 
a barometric condenser 
similar to the type already discussed except for the fact 
that it requires no air pump. 

The barometric or tail pipe of any barometric condenser 
can be replaced by any kind of a pump, and centrifugal 
pumps are often used for this purpose. When large quanti- 
ties of gas are to be handled, as when a dry-air pump 
is not used, the centrifugal pump must be specially 
designed. 




Fig. 175.- 



-Westinghouse-Leblanc Air 
Pump. 



CONDENSERS AND RELATED APPARATUS 267 

A recently developed type of condenser in which the 
barometric tube is replaced by a centrifugal pump and in 
which a separate air pump of a rotary type is used is illus- 
trated in Figs. 175, 176 and 177. It consists essentially of 
the condensing head and well, combined with a centrifugal 
tail pump and a rotary air or vacuum pump as indicated 




Atmosphere 
ReliefValve 






^ Inlet to Well-4— 






V t Submerged noV 1 
fc I less than asja» 



-"Cold Well 

»i%^.-^.^* , s-.*%.S?.«N*«?.-» , ^i«? > j^ 



Fig. 176. — Westinghouse-Leblanc Condenser. 



in Fig. 177. Injection water entering through nozzles in 
the head meets the exhaust, and the resultant mixture 
flows down into the well through the large nozzle shown. 
The liquid is continuously removed from the bottom of this 
well by the centrifugal tail pump and discharged to the hot 
well. The air and associated vapors are drawn down the 
air pipe and discharged by means of the device shown in 



268 



STEAM POWER 



Fig. 175. Water enters the central part of this pump as 
indicated in Fig. 176 and is discharged through the station- 
ary nozzles N and the moving vanes V shown in Fig. 175. 
The water is thus caused to form a series of " pistons " 
which move rapidly downward in the discharge nozzle N f 
and which trap small plugs or lamina of " air " between 

them and thus discharge 

/ 



Exhaust 




the " air " to the atmos- 
phere. The connection 
marked P is used for prim- 
ing at starting when neces- 
sary. 

In small units the cen- 
trifugal tail pump may be 
omitted and the design so 
remodeled that all the injec- 
tion water passes through 
the rotary air pump which 
discharges the entire mix- 
ture from the condenser 
just as it discharges the air 
and associated vapors in 
the larger sizes. 

115. Non-contact Con- 
densers. The type called 
the surface condenser is 
the best-known example of 
non-contact condenser. It 
consists essentially of a large cylindrical or rectangular 
vessel into which the exhaust is discharged and through 
which pass numerous bronze or alloy tubes which carry the 
condensing water, and the surfaces of which act as the 
condensing and cooling surface. 

One form of surface condenser mounted above the 
pumps which serve it is shown in Fig. 178. The exhaust 
enters at the top of the rectangular shell and works its 




Fig. 177. — Westinghouse-Leblanc 
Condenser. 



CONDENSEES AND BELATED APPARATUS 269 







/?" o o o o o ""> 






[ 


x — ' 






^ 


.1 




1 




1 




o ■ ■ ~° 


1111111 


1 




i 


^fP^o^oSoO' 


W&cfifP&JlJ 






O O O O O ^ 







friJH 



270 STEAM POWER 

way down over the water-cooled tubes. The condensate, 
mixed with gases and vapors, is drawn from the bottom of 
the shell by the wet-vacuum pump and discharged to the 
hot well. 

The condensing water is forced through the tubes 
of the condenser by means of the reciprocating circulating 
pump, entering the lower tubes at the right-hand end in 
the figure, making two " passes " through the condenser and 
leaving at the top. Because of the path of the water a 
condenser of this type is sometimes called a two-pass or 
double-flow condenser. 

With the arrangement illustrated, the steam which 
condenses upon the upper tubes falls as a rain from tube 
to tube until it finally settles at the bottom and is drawn 
off. The outer surfaces of the lower tubes are therefore 
practically covered with water and this has two disad- 
vantages. First, these tubes carry the coolest circulating 
water and they therefore cool the condensate coming in 
contact with them while the water flowing through them 
is unnecessarily heated. Cooling of the condensate means 
a lower hot-well temperature than would otherwise be 
obtained, but if the condensate is to be used for boiler 
feed, the temperature of water in the hot well should be 
maintained as high as possible, since this water will eventually 
have to be heated to boiler temperature with a correspond- 
ing expenditure of heat. Second, tubes which are being 
used to cool water covering them are of little use as condens- 
ing surface, and hence a large part of the surface in such a 
condenser is comparatively inactive. 

The ideal arrangement would carry away the liquid 
condensate as fast as formed, leaving the tubes first entered 
by the condensing water to act as the final condensing 
and cooling surfaces, thus bringing gases and non-condens- 
ible vapors into contact with the coolest surfaces just before 
entering the vacuum pump. Numerous designs which 
approximate this ideal have been developed recently and 



CONDENSERS AND RELATED APPARATUS 271 

they give better results than do the earlier and simpler 
forms. The improvement is shown by the values of con- 
densing surface per developed horse-power of engine. In 
early designs it was customary to supply 2 \ sq.ft. of tube 
surface or more per horse-power. Some of the most recent 
installations are giving better vacuums with only 1 sq.ft. 
per horse-power. 

One of the newer models passes the condensate through 
a set of tubes so located that the engine exhaust strikes 
them before impinging on any tubes carrying condensing 
water. This results in a partial condensation of the exhaust 
and raises the temperature of the condensate within the 
tubes to very near that of the exhaust, thus heating the 
boiler feed to a temperature practically corresponding 
to the exhaust temperature of the engine. 

Surface condensers are commonly operated with a 
vacuum of from 24 to 26 ins. of mercury when used with 
reciprocating engines and with a vacuum of 28 to 29 ins. 
when receiving the exhaust of steam turbines. When 
operated at the lower vacuums wet-vacuum pumps are gen- 
erally used, but the best types of dry-air pumps must be 
installed in combination with well-designed condensers 
when the higher vacuums are sought. 

116. Water Required by Contact Condensers. The 
weight of circulating water required varies with the type 
of condenser and with the conditions of operation, such as 
initial temperature of water, vacuum desired, etc. It can be 
determined approximately by calculation and the values 
thus found must then be increased by such factors as 
experience has shown to be necessary. 

In contact condensers the water and the condensate 
are discharged as a mixture and therefore have the same 
average discharge temperature. 

Let h= initial temperature of injection water in F.°; 
h = temperature at which mixture is discharged in 
F.°; 



272 STEAM POWER 

X = total heat above 32° F. of steam as exhausted; 
W = pounds of injection water per pound of exhaust 
steam. 

Assuming the exhaust steam to be dry saturated, each 
pound of steam in condensing to water at a temperature 
of fa degrees must give up an amount of heat equal to X 
minus the heat of the liquid at t°2 or roughly X— (fa — 32) 
B.t.u. This same quantity must be absorbed by the in- 
jection water, while its temperature rises from t\ to fa 
degrees. Each pound of water can then absorb approxi- 
mately {fa — ti) B.t.u. and the pounds of injection water 
per pound of steam will be 

Tr = X-fe+32 

fa — h 

The value of fa would be that corresponding to the 
absolute pressure in the condenser if it were not for the 
air and similar gases which exert some pressure. It is 
generally 10 or more degrees F. below the temperature 
corresponding to the vacuum. Values of fa in the neigh- 
borhood of 110° to 125° F. are customary with recipro- 
cating engines and values as low as 80° are used with high 
vacuums in connection with steam turbines. 

The weight of water used per pound of steam as given 
by Eq. (78) will vary between about 15 for very low initial 
and moderate discharge temperature to about 50 with 
average initial and moderate discharge temperature. Ex- 
perience shows that it is necessary to add 10 per cent or 
more to the values of W obtained from equation (78) to 
obtain the weight of water which will probably be used. 

ILLUSTRATIVE PROBLEM 

Find the quantity of water theoretically required per pound 
of steam condensed in a contact condenser in which a vacuum 
of 25.5 ins. of mercury is maintained when the barometer reads 



CONDENSEES AND RELATED APPARATUS 273 

29.5 ins. of mercury. The initial temperature of the water is 
60° F. 

The absolute pressure in the condenser is 29.5—25.5=4.0 ins. 
of mercury and the steam tables give for this pressure, X = 1115.0 
and to = 126. Substituting in Eq. (78) gives 

Tr 1115.0-126 +32 

T * = ion an =15.5 approximately. 

126—60 

117. Weight of Water Required by Non-contact Con- 
densers. In the case of non-contact condensers there is 
no definite relation between the discharge temperature 
of the cooling water and that of the condensate. Experi- 
ence shows that the discharge temperature of the circulating 
water is generally from 10 to 20 degrees lower than the 
temperature corresponding to the vacuum. 

The temperature of the condensate (hot-well tempera- 
ture) is generally 15 or more degrees below that correspond- 
ing to the vacuum, but good design makes the hot-well 
temperature very closely approximate that corresponding 
to the vacuum. 

Assuming 

ii= initial temperature of injection water in F.°; 

^2 = final temperature of injection water in F.°; 

t c = temperature at which condensate is discharged, i.e., 

hot- well temperature, in F.°; 
X = total heat above 32° F. of steam as exhausted, 

and 

W = pounds of injection water per pound of exhaust 
steam. 

The weight of water which must be circulated per pound 
of steam can be found as in the case of the contact con- 
denser. It is given by 

W = \-U+Z2 

h — h 



274 BTEAM POWER 

Values in the neighborhood of 25 lbs. of water per pound 
of steam are common with low vacuums and 50 or more 
pounds are often used with vacuums over 28 ins. of 
mercury. 

ILLUSTRATIVE PROBLEM 

A surface condenser receives circulating water' at a temper- 
ature of 65° F. and discharges it at a temperature of 80° F. It 
maintains a vacuum of 28.0 ins. with the barometer at 29.5, and 
the temperature of the condensate discharged to the hot well is 
equal to 85° F. Find the quantity of circulating water theoretically 
required. 

This vacuum corresponds to an absolute pressure of 
29.5—28.0 = 1.5 ins. of mercury. Assuming this all due to steam 
(neglecting presence of air) the value of X may be taken from the 
steam table as 1100.1 B.t.u. Substitution in Eq. (79) then gives 

w 1100.1-85+32 ' 

" = ™ — TTr =69.9 approximately. 

80—65 

118. Relative Advantages of Contact and Surface Con- 
densers. The contact types are as a rule much cheaper 
than the surface condensers, and they are less subject to 
serious depreciation, the tubes of surface condensers often 
corroding seriously in very short intervals of time. On 
the other hand, the injection of the cooling water into 
the condensing space in contact types results in the intro- 
duction of large quantities of dissolved gases, and much 
of this material is liberated under the reduced pressure, 
thus tending to increase the condenser pressure, that is, 
decrease the vacuum. Where pumps are used to carry 
away the mixture with contact condensers, these pumps 
have to handle a much larger quantity of water than the 
corresponding pump in a surface condenser, and the work 
of pumping this water out of the vacuum into the atmos- 
phere combined with the additional work required of 
the pump which handles the " air " may partly balance 
the advantage of lower first cost of the contact type. 



CONDENSERS AND RELATED APPARATUS 275 

A surface condenser must always be installed where 
it is desirable to use the condensate as boiler feed, and 
it is generally used when very high vacuums (low absolute 
pressures) are to be maintained. The surface condenser 
is at a serious disadvantage, however, when required to 
handle the exhaust of reciprocating engines. The exhaust 
from such engines always contains large quantities of 
lubricating oil carried out of the cylinder, and unless this 
material is separated before the exhaust enters the con- 
denser it is deposited on the outer surfaces of the tubes 
and decreases the conductivity of those surfaces. Such 
oil can be eliminated to a great extent before the exhaust 
enters the condenser by means of oil separators, which are 
generally made up of a series of baffles upon which the 
steam impinges and upon which the oil is caught. 

119. Cooling Towers. The large quantity of circula- 
ting water required by condensing plants is often an item 
of great , economic importance. When such plants are 
located near a river or near tide water, the circulating 
water can generally be procured for the cost of pumping. 
When they are located in the middle of cities or in regions 
where water is scarce, the cost of water may be excessive 
or it may even be impossible to obtain a continuous supply 
equal to the demand of the condensers. 

In such cases the condensing water is often circulated 
continuously, being cooled after each passage through the 
condensers. This cooling is generally done by exposure 
of a large surface to the air. The resultant evaporation 
of some of the water with the absorption of its latent heat 
of vaporization cools the remainder so that it can be used 
again. This sort of cooling may be effected by running 
the water into a shallow pond of large area, or by spraying 
it into the air over a small pond or reservoir or by passing 
it through a cooling tower. 

Cooling towers are large wood or metal towers generally 
filled with some form of baffling devices. The hot water 



276 STEAM POWER 

is introduced at the top and spread into thin sheets or 
divided up into drops as it descends. Air enters at the 
bottom and flows upward, cooling the water by contact 
and by the partial evaporation which results. The cir- 
culation of air may be natural, i.e., due to the difference 
of temperature between the air inside and out, in which 
case a stack is fitted to the top of the tower; or the cir- 
culation may be forced by fans located in the base of the 
tower. In the latter case the apparatus is called a forced- 
draught cooling tower. 



CHAPTER XV 
COMBUSTION 

120. Definitions. Certain substances are known to 
chemists as compounds, because they can be separated by 
chemical processes into simpler substances. Thus water 
and many of the most familiar materials known to man 
are compounds which can be separated into two or more 
simpler materials. 

Those substances which cannot be further broken up by 
the processes used in separating compounds are called 
elements; they are regarded as elemental, as the stones 
of which the compounds of nature are built up. About 
seventy-five of these elements are now known, but many 
of them are comparatively rare. Pure metals are all 
elements; the oxygen and nitrogen which are mixed to form 
the greater part of the atmosphere are elements; carbon, 
which forms the greater part of most fuels, is an element. 

In many cases the combination of elements to form 
compounds is accompanied by the liberation of heat, and 
some of these combinations are used by the engineer for the 
purpose of obtaining heat in large quantities. When the 
elements which occur in fuels, such as coal, wood and 
petroleum, combine with oxygen, the process is spoken of 
as combustion. The quantity of heat liberated when a 
pound of any material combines with oxygen (burns) is 
called the heat value or calorific value of that material. 

Fuels contain a great number of elements, but only 
three of these ordinarily take part in combustion and are 
therefore spoken of as combustibles. They are carbon, 
hydrogen and sulphur. The sulphur content is generally 

277 



278 STEAM Po ^ER 

very small, and the carbon and hydrogen are therefore the 
most important constituents. 

The combustion of each of these elements will be con- 
sidered in detail in the following sections, but before this 
can be done two other ideas must be developed. 

The smallest particle of an element which can be 
conceived of as entering into combination to form a com- 
pound is known as an atom of that element. It has been 
found that the atoms of each element have an invariable 
and characteristic mass. The lightest atom is that of 
hydrogen, and its weight is considered unity. The atom 
of carbon is twelve times as heavy as that of hydrogen and 
carbon is therefore said to have an atomic weight equal to 
twelve. Similarly the atomic weight of nitrogen is four- 
teen and that of oxygen is sixteen. 

The smallest particle which can be formed by the com- 
bination of atoms is known as a molecule. Like or unlike 
atoms may combine to form molecules. Thus two hydro- 
gen atoms combine to form a molecule of hydrogen, and 
hydrogen gas as it ordinarily exists may be pictured as 
made up of a collection of such molecules. Similarly, 
gaseous oxygen and gaseous nitrogen may be pictured as 
collections of molecules which are made up of two like 
atoms. 

When unlike atoms combine to form a molecule, they 
form a molecule of a compound. Obviously a molecule of 
any compound is the smallest particle of that compound 
which can exist. 

For convenience, the differeDt elements are represented 
by abbreviations; thus oxygen is represented by O, nitro- 
gen by N, hydrogen by H, carbon by C and sulphur by S. 
When these abbreviations are written in chemical equa- 
tions, such as will be given later, they stand for an atom 
of the substance. Hence in a chemical equation would 
mean one atom of oxygen. The symbol O2 is used to mean 
two atoms of oxygen in combination, hence, one molecule 



COMBUSTION 279 

of oxygen. The symbol 2O2 means two groups of two 
oxygen atoms in combination, hence two molecules of oxygen. 
The simplicity and elegance of this system will become 
apparent as the chemical equations which follow are de- 
veloped and explained. 

121. Combustion of Carbon. Carbon can unite with 
ox3 r gen or burn to form two different compounds— carbon 
monoxide (CO) and carbon dioxide (C0 2 ). The monoxide 
is formed by the combination of one atom of oxygen with one 
atom of carbon; the dioxide, by the combination of two 
atoms of oxygen with one of carbon. The dioxide, therefore, 
contains twice as much oxygen as does the monoxide. 

Carbon burned to carbon monoxide has not combined 
with the largest possible quantity of oxygen, and combus- 
tion is therefore said to be incomplete in such cases. When, 
however, carbon dioxide is formed, the carbon has combined 
with as much oxygen as possible and combustion is said to 
be complete. 

It will be shown later that much more heat is liberated 
when the dioxide is formed than when carbon burns to the 
monoxide. Hence, when liberation of heat is the object of 
combustion, the process should be so conducted as to result 
in the formation of the maximum quantity of dioxide and 
the minimum amount of monoxide. 

122. Combustion to CO. The combustion of carbon and 
oxygen to form the monoxide can be represented by the 
equation 

C+0 = CO, (80) 

or by the equation 

2C+0 2 = 2CO (81) 

The former is the simpler and will be considered first, but 
the latter is the more perfect and indicates more to the 
trained eye than does the simpler form. 

The simple equation states that one atom of carbon 
combined with one atom of oxygen to form one molecule 



280 STEAM POWER 

of carbon monoxide. It can, however, be so interpreted 
as to show much more than this. The carbon atom is twelve 
times as heavy as the hydrogen atom, while the oxygen atom 
is sixteen times as heavy as that of hydrogen. The equation 

C + = CO, 

therefore, shows that an atom, which is twelve times heavier 

than the hydrogen atom, unites with one which is sixteen 

times heavier than the hydrogen atom to form a molecule 

which is 28( = 12 + 16) times heavier than the hydrogen 

atom. 

In other words, the weights of carbon and oxygen 

12 3 1 
combining are in the ratio of t^ = j = tT' If a sufficient 

lb 4 I3 

number of carbon atoms to weigh one pound be used, a 
quantity of oxygen weighing 1§ lbs. will be required to 
combine with them to form carbon monoxide. The re- 
sultant carbon monoxide will contain the pound of carbon 
and the 1J lbs. of oxygen and will therefore weigh 2\ lbs. 

The same weight relations would hold irrespective of 
the weight of carbon used, and the simpler equation may 
therefore be put 

1 weight of C+1J weights of = 2J weights of CO. (82) 

ILLUSTRATIVE PROBLEM 

To illustrate the use of this equation, assume that 9 lbs. of 
carbon are burned to carbon monoxide and that it is desired to 
find the weight of oxygen used, and the weight of the product. 
The weight of oxygen used must be 1^ times the weight of carbon, 
that is, lfX9 = 12 lbs. The weight of the product must be 2\ 
times the weight of the carbon, that is 2^X9 =21 lbs.; or, it must 
be the weight of the carbon burned plus the weight of the ox} T gen 
used, that is, 9+12 =21 lbs. 

In general, the oxygen used for combustion is obtained 
from the atmosphere, which may be considered as a median- 



COMBUSTION 281 

ical mixture of oxygen and nitrogen in unvarying porportions. 
These proportions are roughly, 0.23 of oxygen to 0.77 of 
nitrogen by weight, or 0.21 of oxygen to 0.79 of nitrogen 
by volume, as shown in Table VIII. The weight of 
air which contains one pound of oxygen is therefore 

' — - = 4.35 lbs., and this weight of air contains 

U. iu 

4.35-1=3.35 lbs. of nitrogen. 

In the problem previously considered it was found 
that 12 lbs. of oxygen would be required to burn 9 lbs. 
of carbon to CO. The total weight of air required to 
obtain this oxygen will be 12X4.35 = 52.2 lbs. and it will 
contain 52.2 — 12 = 40.2 lbs. of nitrogen. 

By simple arithmetical calculations of the type just 
given all the weight relations involved in the combustion 
of C to CO can be determined. The volume of air required 
in any given case can be found by multiplying the weight 
of air by the specific volume as given in Table VIII. 
Thus, in the illustrative problem already considered, it 
was found that 52.2 lbs. of air would be required to burn 
9 lbs. of C to CO. ' The volume of this air at 62° F. would 
be 52.2X13.14 = 685.9 cu.ft. 

It is found that a quantity of heat equal to about 
4500 B.t.u. is liberated per pound of carbon burned to CO; 
that is the calorific value of C burned to CO is 4500 B.t.u. 

Returning now to Eq. (81), which was said to be more 
useful than the simpler form given as Eq. (80), it will be 
necessary to consider a rather simple law of gases. It 
has been shown experimentally that equal volumes of all 
gases contain the same number of molecules when at the same 
temperature and pressure. This statement is known as 
Avogadro's Law.. It has also been shown that the mole- 
cules of gaseous oxygen contain two atoms. 

The equation in question, 

2C+0 2 = 2CO 



282 



STEAM POWER 



can therefore be read, two atoms of carbon combine with 
one molecule of oxygen to form two molecules of carbon mon- 
oxide. But, if every molecule of O yields two molecules 
of CO it follows from Avagadro's law that the CO formed 
will occupy twice the volume of the oxygen used if measured 
at the same temperature and pressure. If the equation be 
imagined as containing a numeral 1 before the 02, it 
will be obvious that the coefficients of the terms represent- 
ing gas molcules give volume relations directly. This equa- 
tion therefore gives both volume and weight relations. 

TABLE VIII 
Properties of Air 

Considering it to consist only of nitrogen and oxygen. 





Relative Proportions. 


Ratio of N to 0. 


Ratio of Air to O. 




Exact. 


Approx. 


Exact. 


Approx. 


Exact. 


Approx. 


By weight . . 
By volume. . 


/ 0.766 N 
I 0.234 

/ 0.791 N 
10.209 


0.77N 
0.23 O 

0.79 N 
0.21O 


3.27 

3.78 


3.35 
3.76 


4.27 
4.76 


4.35 
4.76 




Spec. wt. at Atmos. Press. 
(Lbs. per Cu.ft.) 


Spec. Vol. at Atmos. Press. 
(Cu.ft. per Lb.) 




At 32° F. 


At 62° F. 


At 32° F. 


At G2° F. 




0.08072 


. 07609 


12.39 


13.14 



Weight of air containing one pound of oxygen, is approximately 
4.35 lbs. 



123. Combustion to CO2. The combination of carbon 
and oxygen to form the dioxide is represented by the equa- 
tion 

C-tr0 2 = C0 2 , (83) 



COMBUSTION • 283 

which shows that one atom of carbon (twelve times heavier 
than hydrogen) combines with two atoms of oxygen (each 
sixteen times heavier than hydrogen) to form a molecule 
of CO2, which is forty-four times heavier than an atom of 
hydrogen. Therefore the weight of carbon and oxygen 

12 3 1 

combining are as 0v/1 =q=^, so that 2| lbs. of oxygen 

ZX-LO o Z% 

are required to burn a pound of carbon to carbon dioxide. 
Writing this in the form of an equation, gives 

1 weight of C+2f weights of = 3| weights of C0 2 . . (84) 

The weight of air required can readily be found by 
multiplying the required oxygen by the number 4.35, 
previously shown to be the number of pounds of air con- 
taining one pound of oxygen. Thus, the required air is 
2|X4.35 = 11.57 pounds per pound of C burned to CO2. 
This number is commonly rounded out to 12 in engineering 
literature. 

The equation given shows volume relations directly. 
It is evident, therefore, that one molecule of O yields one 
molecule of CO2, and hence that the volume of the product 
is exactly equal to the volume of the oxygen used in forming 
it if measured at the same temperature and pressure. 
This is a very important relation, and is often made use 
of in engineering calculations. 

Experiment shows that when carbon burns to the 
dioxide about 14, GOO B.t.u. are liberated per pound of 
carbon burned, that is, the calorific value of C burned to CO2 
in 14,600. 

124. Combustion of CO to CO2. Since carbon which 
has burned to carbon monoxide has not combined with the 
greatest possible quantity of oxygen, the monoxide can 
take up more oxygen by burning to the dioxide. This 
process is represented by the formula 

2CO+0 2 = 2C0 2 , (85) 



284 STEAM POWER 

which shows that two molecules of monoxide combine with 
one molecule of oxygen to form two molecules of the dioxide. 
The volume of CO2 formed is therefore equal to that of 
the CO burned. 

So far as the ultimate result is concerned, it makes no 
difference whether carbon is burned directly to CO2 or is 
first burned to CO and then the CO is burned to CO2. 
The total oxygen used per pound of carbon burned to CO2 
and the total heat liberated per pound of carbon burned 
to CO2 are the same in both cases. 

Thus, for the oxygen, one pound of C burned to CO2 
requires 2f lbs. of oxygen; but one pound of C burned 
to CO requires If lbs. of oxygen, and 1J lbs. more will be 
required when this CO is burned to CO2. The result 
is therefore the same. 

For heat liberated, one pound of C burned to CO2 
liberates about 14,600 B.t.u.; but one pound of C burned 
to CO liberates about 4500 B.t.u. and 10,100 B.t.u. are 
liberated when this CO is burned to CO2. Since the sum 
of 4500 and 10,100 is equal to 14,600 the result is again 
the same. 

Data on the combustion of C to CO and CO2 and the 
combustion of CO to CO2 are collected in convenient 
form in Table IX. 

125. Conditions Determining Formation of CO and CO2. 
Excluding certain complicated considerations which are 
not of great importance in steam-power engineering, it may 
be said that when carbon is being burned at a certain rate 
(pounds per unit of time) the amount of oxygen brought 
into contact with ,the carbon determines whether the carbon 
burns to CO or to CO2. If enough or more than enough 
oxygen to burn the carbon to CO2 is brought into contact, 
that oxide will be formed. If thereMs not enough to burn 
all the carbon to the dioxide, both oxides are formed in cer- 
tain proportions, which can be calculated. 

Since combustion to CO yields only 4500 B.t.u. per 



COMBUSTION 



285 



pound of C and combustion to CO2 yields 14,600 B.t.u. 
per pound of C, the importance of supplying sufficient 
oxygen to burn all carbon to the dioxide in cases where 
the liberation of the maximum quantity of heat is desirable 
is obvious. In actual practice the oxygen is furnished 
by supplying air and it is found necessary in most cases 
to supply more than the amount of air theoretically re- 
quired in order to insure burning all, or even nearly all, 
of the carbon to the dioxide. This comes from the great 
difficulty met in obtaining contact between the oxygen of the 
air and the carbon which is to be burned, that is, in bringing 
all the oxygen of the air into intimate contact with the 
fuel being burned in real apparatus. 



TABLE IX 
Combustion Data for Carbon 

(Per pound of carbon.) 



Product. 


Oxygen Required. 


Nitrogen Accompanying 
Oxygen. 


Pounds. 


Cu.ft. at 62° F. 
and 14.7 Lbs. 


Pounds. 


Cu.ft. at 62° F. 
and 14.7 Lbs. 


CO 


1.333 
2.667 
1.333 


16.0 
32.0 
16.0 


4.46 
8.92 
4.46 


60 1 


C0 2 fromC 

CO, from CO ... . 


120.2 
60.1 





Air Required. 


Quantity of Product 
(N not included). 




Product. 


Pounds. 


Cu.ft. at 
G2° F. and 
14.7 Lbs. 


Pounds. 


Cu.ft at 
62° F. and 
14.7 Lbs. 


Heat Liber- 
ated. 


CO 


5.79 
11.58 

5.79 


76.1 
152.2 

76.1 


2.33 
3.67 

3.67 


32.0 

32.0 

32. cj 


4,500 
14,600 
10,100 per lb. 
of C in CO 
4,300 per 1!). 
of CO 


C0 2 from C 

C0 2 fromCO.... 



286 STEAM POWER 

The air in excess of that theoretically required to burn 
all the carbon completely is spoken of as excess air. In 
the form of an equation, this statement is equivalent to 

Air supplied — air theoretically required = excess air. (86) 

It is customary to express the quantity of excess air in 
terms of a numerical factor known as the excess coefficient. 
This coefficient is defined as the number by ivhich the quantity 
of air theoretically required must be multiplied to give the 
quantity of air actually used. In the form of an equation 
this gives 

Excess coefficient X air theoretically required 

= air actually used. . (87) 

ILLUSTRATIVE PROBLEM 

Taking data from the illustrative problem previously considered, 
assume that 9 lbs. of carbon are burned in air to C0 2 . Each pound 
theoretically requires 11.57 lbs. of air, so that the theoretical 
air-supply for this case would be 9X11.57=104.13 lbs. If in a 
real case 150 lbs. of air are supplied, the excess coefficient is equal 
to 150+- 104.13 =1.44. 

126. Flue Gases from Combustion of Carbon. The 

gases resulting from the combustion of fuels are known 
in engineering as the products of combustion or flue gases, 
because they are the gases passing through the flues or 
passages leading from furnaces in which fuel is burned and 
to the stacks which serve to carry off the gases. 

It has already been shown that the CO2 formed by the 
combustion of carbon has the same volume as the oxygen 
which is used in forming it. Therefore, if the air supplied 
in a given case just equaled that theoretically required 
for combustion to CO2 and if all of the oxygen were used, 
the CO2 formed would merely replace the oxygen in the 
air. The theoretical proportions of the flue gas would 
then be 0.21 of C0 2 and 0.79 of N by volume. 



COMBUSTION 



287 



If real flue gases obtained by burning carbon in air are 
found to contain less than 21 per cent of CO2, the combustion 
has evidently not yielded theoretically perfect flue gases. 
The trouble may be due to an excess or to a deficiency of air. 
If there is an excess of air there will be oxygen present in 
the flue gases; if there is a deficiency there will be CO 
present in the flue gases. An analysis of these gases for 
oxygen and for CO would therefore indicate the source of 
trouble and the remedy to be provided. 

The curve to the right of the central vertical line in 
Fig. 179 shows the theoretical decrease in volume per 



35 

©30 

s 

>>20 



8l5 



« 10 

o 

£5 



\ 






















\ 
\ 














































\ 

\ 


\ 






















^\ 




















& 




\ 








% 
















\ 















50 40 



20 10 



Deficiency (in per cent) 



50 100 150 200 250 300 
Excess (in per cent) 

2 3 4 



Fig. 179.— Effect of Air Supply on Flue Gas Analysis. 






cent of CO2 in flue gases as the excess air increases. The 
single numbers 1, 2, 3 and 4 indicate the excess coeffi- 
cients corresponding to the various percentages of excess 
air. 

The curves to the left give the theoretical decrease in 
volume per cent of CO2 and the theoretical increase in 
volume per cent of CO as the air supplied is decreased below 
that theoretically required for complete combustion. 

127. Combustion of Hydrogen. Hydrogen combines 
with oxygen, or burns, to form water. The equation for 
this reaction is 

2H 2 +0 2 = 2H 2 0, (88) 



288 STEAM POWER 

which indicates that two molecules of hydrogen combine 
with one molecule of oxygen to form two molecules of 
water. In terms of volumes, two volumes of hydrogen 
combine with one of oxygen to form two of gaseous water, 
that is, water in the form of highly superheated vapor. 
As the water is cooled down it will obviously approach 
and finally reach the liquid condition, with a rapid de- 
crease in volume quite different from that experienced 
by a gas under similar conditions, so that the volume rela- 
tions hold only at high temperatures. 

The weight relations can be calculated as in other 
cases, starting from the fact that four weights of hydrogen 
combine with thirty-two weights of oxygen to form 36 
weights of water. The weights of hydrogen and oxygen 
are therefore in the relation of ^ = |. 

The heat liberated when one pound of hydrogen burns 
to water is equal to about 62,000 B.t.u. This is the quantity 
of heat which could be obtained if one pound of hydrogen 
at, say, room temperature, and mixed with the requisite 
quantity of oxygen, were ignited and the resultant water 
were then cooled down to the initial temperature. During 
the cooling of the water it would partly or entirely condense 
and thus give up some or all of its latent heat of vaporization. 
This heat would obviously be included in the calorific 
value just given. 

In many pieces of engineering apparatus in which 
hydrogen is burned the products of combustion are not 
cooled to such an extent that the water is condensed. The 
latent heat of vaporization would not be liberated under 
such conditions, but would remain bound up with the water 
vapor. When the water is not condensed the heat liberated 
is only about 52,000 B.t.u. per pound of hydrogen. This 
number is known as the lower calorific value of hydrogen, 
while 62,000 is known as the higher calorific value. 

Data on the combustion of hydrogen are given in 
Table X. 



COMBUSTION 



289 



TABLE X 

Combustion Data for Hydrogen 

(Per pound of hydrogen) 





Oxygen Required. 


Nitrogen Accompanying Oxygen. 


Product. 


Pounds. 


Cu.ft at 62° F. 
and 14.7 Lbs. 


Pounds. 


Cu.ft. at 62° F 
and 14.7 Lbs. 


H 2 


8 


96 


26.8 


361 





Air Required. 


Quantity of Product (N 
not included). 




Product. 


Pounds. 


Cu.ft. at 62° 

F. and 14.7 

Lbs. 


Pounds. 


Cu.ft. at 62° 

F. and 14.7 

Lbs. 


Liberated. 


H 2 


34.8 


457 


9 


Liquid 
0.144 


/ 62,000 
\ 52,000 



128. Combustion of Hydrocarbons. Many of the fuels 
used by the engineer contain compounds of hydrogen 
and carbon which are called hydrocarbons. One of the best 
examples is methane (CH4), which forms the greater part 
of all the so-called natural gas. 

All of these hydrocarbons burn to CO2 and H2O if 
the supply of oxygen is great enough. If there is a deficiency 
of oxygen, combustion is incomplete and generally results 
in the formation of CO2, H2O, CO, C in the form of soot, 
and other products which need not be considered here. 

For complete combustion the requisite oxygen and 
air can be determined as in previous cases by means of 
chemical equations. Thus for methane the equation is 



CH4+20 2 = C02+2H 2 0, 



(89) 



which shows that sixteen (12+4) weights of methane 
combine with sixty-four (2X2X16) weights of oxygen to 
form forty-four (12+32) weights of carbon dioxide and 
thirty-six (4+32) weights of water. 



290 STEAM POWER 

The calorific value of hydrocarbons is generally assumed 
to be equal to the sum of the heat values of the carbon 
and hydrogen contained in one pound of the material. 
Thus, if C represent the fraction of a pound of carbon 
contained in one pound of the hydrocarbon and if H 
represent the fraction of a pound of hydrogen contained 
therein, the common assumption would make the higher 
calorific value of the hydrocarbon 

(CX 14,600) + (HX 62,000) B.t.u. . . (90) 

The results obtained in this way do not generally check 
well with the experimentally determined values, and it is 
best to use the latter when they are available. 

129. Combustion of Sulphur. Sulphur forms several 
different oxides, but when burned under engineering con- 
ditions it is generally assumed to form only the dioxide 
SO2. The chemical equation for such combustion is 

S+0 2 = S0 2 , (91) 

and since the atomic weight of sulphur is 32, this equation 
shows that equal weights of sulphur and oxygen combine 
to form the dioxide. 

The combustion of sulphur to SO2 liberates about 
4000 B.t.u. per pound of sulphur. 

130. Combustion of Mixtures. It is often necessary 
to obtain approximate calorific values of combustible 
materials which, without great error, can be considered 
as mixtures of combustible and non-combustible elements. 
If there is oxygen present in the mixture it is assumed 
to be combined with hydrogen in the form of water, so 
that the uncombined or available hydrogen per pound of 
material is given by the expression 

Available H = H-^, (92) 

8 



COMBUSTION 291 

in which H and respectively represent the fractions of 
a pound of hydrogen and oxygen in one pound of material. 
The calorific values of such a mixture containing car- 
bon, hydrogen and sulphur would then be given approxi- 
mately by the equation 

Higher B.t.u. = 14,6000+62,000 (H-^ J +4000S, (93) 

in which the letters stand respectively for the fractions 
of a pound of each of the elements present in one pound of 
the mixture. Similarly the lower calorific value would 
be (approximately) 

Lower B.t.u. = 14,600C +52,000 /ll-^ +4000S, (94) 
and the oxygen required will be 

Pounds of = 2§C+8(h-^\+S. - . (95) 



131. Temperature of Combustion. If combustion of 
any material could be carried on inside of an ideal vessel 
which did not absorb nor transmit heat, the heat liberated 
during the combustion could not escape from the space 
within the vessel. 

If the vessel contained initially only the combustible 
and the oxygen or air required to burn it, the products of 
combustion would be the only material contained within 
the vessel after the completion of combustion. Under 
such circumstances the heat would be used in raising the 
temperature of the products of combustion, and the process 
could be pictured as though all of the combustion occurred 
first, forming the products of combustion without change 
of temperature, and then the liberated heat raised the 
temperature of these products. 

Knowing the weight of each of these products and the 
quantity of heat required to raise the temperature of one 



292 • STEAM POWER 

pound of each of them one degree, the amount of heat 
required to raise all of them one degree could be found by 
multiplying the two known values. Thus, if carbon had 
been burned in oxygen to CO2 with the theoretical oxygen 
supply, the vessel would contain only carbon dioxide. 
To raise the temperature of one pound of carbon dioxide 
one degree requires an amount of heat equal to the specific 
heat of that gas. Therefore, if W represents the weight 
of CO2 formed and C represents its specific heat, the amount 
of heat required to raise the temperature of all of the CO2 
one degree would be W-C B.t.u. If Q B.t.u. were liberated 
by the combustion, the temperature rise in degrees would 
therefore be given by 

Temp, rise = Sr„ (96) 

and if the initial temperature had been to degrees, the final 
temperature would be 

^'o+Wc {97) 

A final temperature figured in this way is called the theo- 
retical temperature of combustion. It can never be attained 
in practice because of heat lost to surroundings and because 
of other losses which need not be considered here. 

Theoretical temperatures of combustion are, moreover, 
nearly always calculated on the assumption that the specific 
heats of gases are constants, whereas they really increase 
with the temperature. It therefore follows that tempera- 
tures determined on the assumption of constant specific 
heat will be too high for this reason also. 

When gases are heated there are two distinctly different 
limiting possibilities; the volume occupied by the gases 
may remain constant or the pressure exerted by the gases 
may remain constant while the volume increases. In 
the case of constant volume all the heat added to the gases 



COMBUSTION 293 

must be used for raising the temperature; the amount 
of heat required per pound per degree under these con- 
ditions is known as the specific heat at constant volume 
and is designated by C v . 

When, however, the volume is allowed to increase at 
such a rate as to keep the pressure constant the heat sup- 
plied must not only raise the temperature, but must also 
do whatever external work is done in displacing (pushing 
out of the way) surrounding mediums. The heat required 
per pound per degree under these conditions is known as 
the specific heat at constant pressure and is represented 
by Cp. It is always greater than C v by the amount of heat 
required to do the external work accompanying a rise of 
temperature of one degree. 

Thus, in the case assumed above, had the vessel been 
so constructed that its internal volume did not change, 
the specific heat at constant volume would be used. On 
the other hand, had the vessel been fitted with a movable 
piston arranged to move outward at such a rate as to main- 
tain constant pressure within the vessel as the temperature 
rose, the specific heat at constant pressure would be 
used. 

In most cases the material burned is not pure carbon, 
but a fuel containing carbon, hydrogen and sulphur, and as 
air is generally used to furnish the oxygen, the products 
of combustion will contain not only the oxides of carbon, 
hydrogen and sulphur, but inert nitrogen as well. The 
temperature rise is determined in the same way, however, 
by dividing the heat liberated by the amount required to 
raise the temperature of the products one degree. Thus 
if Wij W2, W3 . . . W n stand for the weights of the various 
products and Ci, C2, C3 . . . C n for their respective specific 
heats, the theoretical temperature rise is given by 

Tcmp - rise = ir 1 C 1 +^C 2 +TO+ ■ . . +W.C. ' (98) 



294 STEAM POWER 

and the theoretical temperature of combustion is given by 

t==t0 + W 1 C 1 + W 2 C 2 + WsC3 . . . W u Cn' ^ 

if to stands for the initial temperature. 

PROBLEMS 

1. Assume 10 lbs. of C burned to CO. Determine the quan- 
tity of oxygen required, the quantity of air required, the quantity 
of nitrogen in this air, and the quantity of heat liberated. 

2. What will be the volume of the CO formed as above if 
measured at G2° F. and 14.7 lbs. pressure? 

3. Assume 15 lbs. of C burned to C0 2 . Determine the quan- 
tities of oxygen and air required, the quantity of nitrogen con- 
tained in this air, and the quantity of heat liberated. 

4. What will be the volume of the C0 2 formed from 15 lbs. 
of carbon if measured at 62° F. and 14.7 lbs.? 

5. What will be the volume of the flue gases formed by the 
combustion of 11 lbs. of carbon to C0 2 with the theoretical air 
supply? 

6. The quantity of CO obtained by the combustion of 8 lbs. 
of carbon is burned to C0 2 with the theoretical amount of oxygen. 
Determine the quantities of oxygen and air required, the amount 
of nitrogen contained in this air, and the quantity of heat liberated. 

7. Assume 5 lbs. of C burned in air to C0 2 with an excess 
coefficient of 1.5. Determine the quantities of oxygen and air 
supplied, the heat liberated and the composition of the flue gases. 

8. The composition of flue gases resulting from the combus- 
tion of carbon in air is found to be 21% of C0 2 and 79% of N 
by volume. What is the value of the excess coefficient? 

9. An analysis of flue gases resulting from the combustion 
of carbon in air shows 12% of C0 2 by volume and no CO. The 
gases are not analyzed for or N. What can j^ou say with regard 
to the air supply? 

10. Three pounds of hydrogen burn with theoretical oxygen 
supply. Determine the weight of oxygen and air used, the weight 
of the resultant water and the weight of the flue gas. 

11. Determine the heat liberated in the preceding problem if 
the water vapor is condensed and if it is not condensed. 

12. How much hydrogen would have to be burned to obtain 
20 lbs. of water? 



COMBUSTION 295 

13. The chemical formula of methane is CH 4 . If one pound 
of methane is burned with theoretical air supply, what weight 
of air will be used, and what will be the weight of the flue gases? 

14. What would be the percentage composition of the flue 
gases of the preceding problem on a weight basis? 

15. The chemical formula of ethane is C 2 H 6 . Determine the 
calorific value of this material by means of the formula given 
in the text. 

16. A certain material is found to have the following analysis 
on a weight basis: C, 85%; H, 12%; O, 1%; S, 2%. Determine 
the calorific value of this material by means of the formula given 
in the text, assuming that all the oxygen present is in combination 
with hydrogen. 

17. Determine the amount of oxygen required to completely 
burn 3 lbs. of the material described in the preceding problem. 

18. One pound of carbon is burned to C0 2 in pure oxygen 
in a vessel so arranged as to maintain constant internal pressure. 
The specific heat of C0 2 at constant pressure and at ordinary 
temperatures is about 0.21. Calculate the theoretical temperature 
rise and the temperature of combustion, using this value of the 
specific heat and assuming an initial temperature of 60° F. 

19. Make the same calculations as called for in the preceding 
problem, but using the value 0.27 for the specific heat of C0 2 . 
This is more nearly the average value of the specific heat over 
the range of temperature existing in such a case. 

20.- The hydrocarbon ethylene is represented by the chemical 
formula C 2 H 4 . Assume that one pound of this material is burned 
in air within a vessel arranged to maintain the products at con- 
stant pressure and that the excess coefficient is 1.5. Determine 
the theoretical temperature of combustion if the initial temperature 
is 60° F., the mean specific heat of C0 2 is 0.27, that of H 2 is 
61, that of N is 0,27, and that of O is 0.24. 



CHAPTER XVI 

FUELS 

132. Commercial Fuels. In engineering practice any- 
thing which is combustible and which can be procured in 
large quantities at a reasonable cost is called a fuel. The 
principal commercial fuels are: 

f(l) Coal. 
a. Solid j (2) Wood and wood wastes. 
1(3) Vegetable wastes. 

{(1) Crude petroleum or natural oil. 
(2) Various products made from petroleum. 
(3) Methyl and ethyl alcohol. 

(1) Natural gas. 

(2) Artificial or manufactured gases. 



c. Gaseous 



Coal is by far the most extensively used fuel because of 
its abundance and relative cheapness in most localities. 
However, in oil-producing regions the crude oil and some 
of the products made from it are more often the commonly 
used fuel, particularly if good coal is not mined in the 
immediate vicinity. 

Wood is, in general, too valuable to be used exclusively 
as a fuel excepting on the frontiers where wooded terri- 
tory is being opened up and where coal cannot be pro- 
cured excepting at prohibitively high cost. Wood wastes, 
on the other hand, are very often used for fuel in the indus- 
tries producing them. 

Vegetable wastes, like wood wastes, are essentially of 
local value, being practically entirely consumed by the 
industries producing them. 

296 



FUELS 297 

Natural gas is in many respects an ideal fuel, and is 
extensively used for power production in some localities. 
The diminution in the quantity available, the consequent 
rise in the price, the great economy achieved in burning 
this gas in gas engines and the increased use of the gas for 
domestic purposes are, however, gradually eliminating this 
fuel from the steam-power field. 

Artificial gases have never been extensively used for the 
generation of steam, as it is generally cheaper to burn the 
materials from which the gases are made, rather than to 
convert them into gas and then to burn the gas under 
boilers. This condition may change in the future when 
better markets have been opened up for some of the by- 
products which can be obtained from artificial gas plants. 

133. Coal. The word coal is used as the name of a 
great group of natural fuels which consist of more or 
less metamorphosed vegetable remains. At one end of the 
group is the material known as peat, which is only slightly 
changed from the original vegetable substance; at the 
other end is the graphitic anthracite which has undergone 
such radical metamorphosis that practically all of the 
original vegetable material excepting carbon and ash has 
been eliminated. 

A common, rough classification of the coals in the order 
of age, or of completeness of carbonization is, 

1. Peat or turf. 

2. Lignite (brown or black). 

3. Sub-bituminous coal. 

4. Bituminous coal. 

5. Semi-bituminous coal. 

6. Semi-anthracite. 

7. Anthracite. 

8. Graphitic anthracite. 

The divisions are not at all exact, as they depend partly upon 
chemical composition and partly upon physical properties. 



298 



STEAM POWER 



Another classification of a more exact variety is that 
given in Table XI and partly illustrated in Fig. 180, which 
gives what is known as Mahler's curve. It is for United 
States coals only. The terms used in this classification 
are explained in subsequent paragraphs. 



15000 




70 80 90 

# Fixed Carbon in the Combustible 



100 



Fig. 180.— Mahler's Curve for United States Coals. 



TABLE XI 
Classification of Coals 



Division. 


Per cent of 

Fixed Carbon 

in Combustible. 


Per cent of 
Volatile Matter 
in Combustible. 


Calorific Value, 

B.t.u. per Pound of 

Combustible. 


Graphitic 


100 to 97 
97 to 92.5 
92.5 to 87.5 
87.5 to 75 
75 to 60 
65 to 50 
under 50 


to 3 

3 to 7.5 

7.5 to 12.5 

12.5 to 25 

25 to 40 

35 to 50 

over 50 


14,600 to 14,900 
14,900 to 15,300 
15,300 to 15,600 
15,600 to 15,900 
15,800 to 14,800 
15,200 to 13,700 
13,700 to 11,000 


Anthracite. . 


Semi-anthracite 

Semi-bituminous 

Eastern bituminous. . . 
Western bituminous. . 
Lignite 





The graphitic anthracite occurs in very small quantities 
and mostly in Rhode Island. With a few minor exceptions 
the anthracites occur only in Eastern Pennsylvania and the 



FUELS 299 

semi-anthracites are almost entirely confined to the western 
edge of this field. 

The semi-bituminous coals are found on parts of the 
eastern border of what is known as the Appalachian coal 
field, extending from central Pennsylvania through the 
intermediate States to the northern part of Alabama. The 
greater part of this enormous bed consists of eastern bitu- 
minous coal. Western bituminous coals are found in large 
beds in the central part of the United States, principally in 
the States of Illinois, Indiana and Kentucky on the east of 
the Mississippi River, and in Iowa, Kansas and Texas to 
the west of that river. 

Lignite is found in small quantities in nearly all of the 
western half of the United States and in large beds in the 
Dakotas, Texas, Arkansas, Louisiana, Mississippi and 
Alabama. 

Peat is distributed in small beds throughout practically 
all of the United States and is continually forming in many 
marshes and on low-lying lands. 

134. Coal Analyses. Two different coal analyses are 
in use, the simpler being known as the proximate analysis 
and the more exhaustive being called the ultimate analysis. 
Both are made and reported on a weight basis. 

The proximate analysis assumes coal to contain four 
different and separable things, which are called fixed carbon, 
volatile hydrocarbon or volatile matter or volatile, moisture 
and ash. 

Moisture is determined by maintaining a small quantity 
of finely ground coal at a temperature of about 220° F. 
for one hour. The material lost during this time is 
assumed to be moisture only and is reported as such. 
Coal from which the moisture has been driven in this way 
is called dry coal. 

Volatile matter is determined by heating a sample 
from which the moisture has been driven, or a fresh sample. 
The coal is maintained at a red to white heat with exclu- 



300 STEAM POWER 

sion of air until there is no further loss of weight. In 
the case of a previously dried sample the loss under these 
conditions is called volatile hydrocarbon. If the sample 
was not previously dried a separate moisture determina- 
tion is made on a similar sample and the weight of volatile 
is found by difference. 

Fixed carbon is found by combustion of a sample from 
which the moisture and volatile have been driven, the 
loss under these conditions being assumed to be entirely 
due to the combustion of carbon. 

Ash is the name given to the incombustible material 
left behind after determining the fixed carbon. 

The volatile hydrocarbons and the fixed carbon as 
determined in the proximate analysis are assumed to be 
the only combustible parts of the coal and their sum is 
called the combustible. 

Proximate analyses are reported in three different 
ways: On coal as received, on dry coal, and on combustible. 

Since the water content of a sample of coal received 
at any plant is largely a matter of the weather conditions 
during shipment, the best idea of the character of a coal 
can be obtained by excluding the consideration of its 
moisture content. It is generally best, therefore, to convert 
analyses to a dry coal basis, that is, recalculate the per- 
centages of volatile, fixed carbon and ash on the assumption 
that the analysis was made on the weight of coal which would 
result from drying the sample that was actually used. Ex- 
cessive moisture is, however, undesirable for steam-raising 
purposes, and the amount of moisture should therefore be 
determined in every case. 

Ash is also more or less a matter of accident in that the 
amount contained is largely determined by the care used 
in mining and subsequent cleaning of the coal. While 
it has a very appreciable effect upon the character of the 
material as a fuel it really has little connection with the 
combustible part of the fuel. For purposes of classifica- 



FUELS 301 

tion, therefore, the ash should also be eliminated and the 
analysis given on the basis of combustible. 

Sulphur is sometimes reported with a proximate analy- 
sis. In making such an analysis the greater part of the 
sulphur is really driven off with, and regarded as, part of 
the volatile, so that when the sulphur content is desired it 
must be determined by a separate analysis. 

The ultimate analysis attempts to separate the dry 
combustible into the various elements of which it is com- 
posed. The percentages of carbon, hydrogen, oxygen, 
nitrogen and sulphur are determined as well as the per- 
centage of ash in dry coal. Such analyses show the carbon 
contents of coal to vary from about 98 per cent in the 
graphitic anthracite through about 97 per cent in the 
semi-anthracite, 87 per cent in semi-bituminous, 80 per cent 
in bituminous and 74 per cent in lignites to as low as 61 
per cent in peats. The corresponding figures for hydrogen 
run from about 1 per cent through a range in the neigh- 
borhood of 5 per cent for semi-bituminous to about 6 per 
cent in the case of peat. 

Oxygen varies from about 2 per cent or less in the 
case of good anthracite to as high as 33 per cent for peat; 
nitrogen generally forms about 1 per cent of the dry fuel 
and sulphur from 1 to 3 per cent. 

135. Calorific Value of Coals. The calorific value of 
coals on a basis of combustible has been shown to vary 
approximately according to a smooth curve, but the local 
variations are so great that no generally applicable formula 
for calorific value has yet been proposed. The formula 
commonly used is based upon the ultimate analysis and 
is similar to that suggested as approximately applicable 
in the case of mixtures of combustibles. It is known as 
Dulong's formula, and is 

[62,000 
B.t.u. perlb. = 14,600C + or \( H-^ )+4000S, (100) 

[52 ; 000 




302 STEAM POWER 

in which the letters refer to the weight of the various ele- 
ments contained in one pound of dry coal. 

When an accurate knowledge of the calorific value of a 
fuel is desired it should be obtained by means of a fuel 
calorimeter. There are many varieties of this instrument, 
but practically all operate on the same general principle. 
A known weight of fuel is completely burned within a vessel 
and the heat liberated is absorbed by water or similar 
liquid. From measurements of liquid temperatures the 
heat absorbed by the liquid can be determined, and this 
with some additions for losses of various kinds must be the 
heat liberated by the fuel. 

136. Purchase of Coal on Analysis. Until quite recently 
it was customary to buy coal from the lowest bidder pro- 
vided the material supplied could be made to give satis- 
factory results in the plant. Obviously the purchaser knew 
nothing regarding his purchase, and often bought quantities 
of ash and moisture at the price of combustible. Now, 
however, the larger power plants and many of the smaller 
are buying on the basis of analyses and calorific values as 
determined in calorimeters. 

A certain desirable standard analysis is set and cer- 
tain variations are allowed from it. Wide variations are 
penalized by deducting so many cents per ton for each 
variation of a certain degree, and, finally, outside limits 
are set for moisture and ash beyond which the fuel need not 
be accepted. In some cases limits are also set for sulphur. 

This is the logical method of purchasing coal in large 
quantities, and is sure to come into very general use as its 
advantages become known. 

137. Petroleum. This material is obtained from drilled 
wells and has been found in many widely separated sections 
of the country. The oil wells of Pennsylvania and neigh- 
boring States, of Oklahoma, Texas and California have been 
the most productive and are hence the most widely known. 

Natural petroleum, as it occurs in the United States, is 



FUELS 



303 



generally a dark, rather thick, oily liquid with a char- 
acteristic odor. It varies widely in composition so far as 
the compounds contained are concerned, but the variations 
in ultimate composition, specific gravity and calorific value 
are comparatively small. 

The ultimate analysis of crude oil generally shows about 
83 to 85 per cent of carbon, 13 to 15 per cent of hydrogen 
and small quantities of oxygen, nitrogen and sulphur. 

The specific gravity generally lies between 0.80 and 0.90 
and in most cases is nearer the upper figure. It is common 
practice to express the gravity in terms of the Beaume 
scale, an arbitrary scale developed for an instrument known 
as the Beaume hydrometer. This device is arranged to 
float in liquids and measures the gravity by the distance to 
which it sinks. Various corresponding values of the Beaume 
scale and specific gravity are given in Table XII for the 
region most used in connection with petroleum. 

TABLE XII 
Corresponding Beaume Readings and Specific Gravities 



Beaum6 Reading. 


Specific Gravity. 


Beaume" Reading. 


Specific Gravity. 


20 


0.9333 


34 


. 8536 


22 


0.9210 


36 


0.8433 


24 


0.9090 


38 


0.8333 


26 


0.8974 


40 


0.8235 


28 


0.8860 


42 


0.8139 


30 


0.8750 


44 


0.8045 


32 


0.8641 


46 


. 7954 



The higher calorific value varies between 19,000 and 
20,000 B.t.u. per pound and the lower value is generally 
1000 to 1500 B.t.u. lower. 

Crude oil is sometimes used for fuel, but this is unde- 
sirable, for two reasons. First, the crude oil contains 
many highly volatile constituents which can be distilled 



304 STEAM POWER 

off and which have a high market value in the forms of 
gasoline and allied distillates. Second, the presence of 
these highly volatile constituents in the oil makes it more 
dangerous, as combustible vapors are given off in large quan- 
tities at low temperatures and the mixtures formed with 
the oxygen of the air are often highly explosive. 

As a consequence, the material generally sold as fuel 
oil is a residuum left after distilling off the more volatile 
constituents of the crude oil. It has practically the same 
properties as the crude, excepting that dangerous vapors 
are not given off at so low a temperature. 

PROBLEMS 

1. A sample of coal gives the following proximate analysis: 
moisture, 5%; volatile, 4.25%; fixed carbon, 80.75%; and 
ash, 10%. Determine the percentage of combustible and the 
percentages of fixed carbon and of volatile in the combustible. 

2. What variety of coal is indicated by the values obtained 
in Prob. 1? 

3. The following results were obtained in making a proximate 
analysis of a sample of coal; moisture, 7%; fixed carbon, 56.7%; 
volatile, 24.3%; ash, 12%. Determine the percentage of com- 
bustible and the percentages of fixed carbon and of volatile in the 
combustible. What variety of coal is indicated by these values? 

4. The ultimate analysis of a sample of dry coal gave the 
following results: carbon, 79.12%; hydrogen, 4.14%; oxygen, 
1.84%; sulphur, 0.92%; nitrogen, 0.74%; ash, 13.24%. Recal- 
culate these values for an ash-free coal. 

5. Determine by means of Dulong's formula the upper and 
lower calorific values of the coal described in Prob. 4. 

6. The ultimate analysis of a sample of crude petroleum from 
which all water was removed gave the following results: carbon, 
85%; hydrogen, 13%; sulphur, 1.0%; oxygen, 0.25%; nitrogen, 
0.12%; ash (sand and similar material), 0.63%. Determine the 
upper and lower calorific values by means of Dulong's formula, 



CHAPTER XVII 

STEAM BOILERS 

138. Definitions and Classification. The term boiler 
is generally applied to the combination of a furnace in which 
fuel may be burned continuously and a closed vessel in which 
steam is generated from water by the heat liberated within 
the furnace. 

Boilers are classified in many different ways, the more 
important being given in the following schedule: 

Classification of Boilers 



(1) According to form 



(a) Plain cylindrical, 

(b) Flue, 

(c) Tubular, 

(d) Sectional, etc. 



(2) According to location off (a) Externally fired, and 
furnace } (b) Internally fired. 



(3) According to use 



(a) Stationary, 

(6) Portable (as on trucks, 

or rollers), 
(c) Locomotive, 
1(d) Marine. 



,,v A j. ,. ,. j.f(a) Horizontal, 

(4) According to direction of ' T ,. , 

. . , . \(b) Inclined, 

principal axis \ ' Tr ,. . 

1(c) Vertical. 

(5) According to relative posi-f Water tube 

tions of water and hot{ „: ^. , . ' 

(b) Fire tube, 
gases I 

305 



306 



STEAM POWER 



Examples of boilers of the different types mentioned are 
given in subsequent paragraphs. 

139. Functions of Parts. It has been shown that there 
are two essentially different parts in the apparatus commonly 
known as a steam boiler, the furnace and the boiling vessel. 
A simple form of boiler known as a horizontal, return tubu- 
lar boiler, or an H.R.T. boiler, is shown in Figs. 181 and 
182 with the two essential parts and their components 

Pressure Regulator 



■ .""J 



Steam 
Dome 

m/////////////// /////////M 




Fig. 181.— Sectional Elevation of H.R.T. Boiler and Furnace. 



indicated. The furnace consists essentially of the combina- 
tion of grates, bridge wall, fire and ash doors, the ash pit 
and the space above the grates. It is the function of the 
furnace to so burn the fuel that the maximum amount 
of heat will be made available for absorption by the water 
within the boiling vessel. 

It is the function of the boiling vessel to transmit to 
the water within it the greatest possible quantity of the 
heat thus made available and to resist successfully the 



STEAM BOILERS* 



307 




tendency to rupture under the action of the high internal 
pressure, that is, the pressure of the steam. 

In the type of boiler shown the fuel is " fired " by 
hand, that is, it is 
spread on the grate by 
being thrown from a 
scoop shovel through 
the opened fire door. 
Air enters through both 
doors in regulated pro- 
portions and in such 
quantities as best to 
approximate complete 
combustion. 

The hot gases re- 
sulting from the com- 
bustion pass over the 
bridge wall, along the Fig. 182. 

lower part of the boiler Section through Furnace of H.R.T. Boiler, 
shell and then through 

the fire tubes, or flues, toward the front of the boiler as 
shown by arrows in the figure. From the front end of the 
tubes the products of combustion pass up through the 
smoke box to " breechings " or " flues," which carry them 
to the stack. 

Heat is received by the water within the vessel in two 
different ways: 

(1) The hot fuel bed on the grate radiates energy in the 
same way that the sun or any other glowing body radiates 
energy. Some of this energy traverses the space between 
fuel bed and boiler shell and ultimately passes through that 
shell to the water within. The rest of the radiated energy 
passes into the walls surrounding the furnace and heats 
them and the surrounding atmosphere. 

(2) The hot gases of combustion pass over the heating 
surface of the boiler, as shown, and transmit part of their 



308 STEAM POWER 

heat to the water on the other side of those surfaces. The 
rest of the heat which they carry is either lost to the surround- 
ing walls or is carried up the stack by the gases which 
leave the boiler at a comparatively high temperature. 
This temperature ordinarily ranges from about 500° to 700° 
F. and in extreme cases goes even higher. 

140. Furnaces and Combustion. In most forms of 
boiler the water within the boiler has practically the same 
temperature as- the steam being generated, and this is 
generally from 320° to 400° F. Obviously the products 
of combustion cannot be cooled by the water to a tem- 
perature below that of the water, so that the gases leaving 
the boiler in an ideal case would have a comparatively 
high temperature. Practically, it is found undesirable to 
attempt to reduce the temperature of the gases to a value 
even approximating that of the water and, as indicated 
above, they are discharged at a temperature several hundred 
degrees higher. In order that the maximum amount of 
heat may be made available for the boiling vessel the prod- 
ucts of combustion must therefore leave the furnace with 
the highest possible temperature, and the ideal furnace 
would completely burn the cheapest fuel available in such 
a way as to give this highest possible temperature and not 
to generate smoke. 

Real furnaces fall far short of this ideal performance, 
for numerous reasons. The more important of these are 
given in the following paiagraphs: 

(a) Complete Combustion of Carbon. In a real furnace 
the combustion of the carbon of the fuel may be incomplete 
in two senses; first, some of the carbon may remain entirely 
unoxidized and pass off with the ash, and second, some of 
the carbon may be burned to CO instead of to CO2. 

Imperfect combustion of the first kind can result from 
fuel falling through the openings in the grate before it has 
been ignited or when only partly burned, or it can result 
from failure to get air to some of the carbon in sufficient 



STEAM BOILEES 309 

quantities to burn it completely before all of the surrounding 
fuel has been converted into ash and the locality cooled 
down to such an extent as to allow the unburned carbon 
in its midst to cool below the temperature of ignition. 

Imperfect combustion of the second kind, resulting 
in the formation of CO, generally results either from a 
lack of sufficient air above the fuel bed or from an excessive 
quantity of air above the bed. In any furnace there is a 
tendency toward the formation of CO within the bed of fuel, 
and the deeper the bed the greater this tendency. If the 
CO thus formed meets sufficient air after leaving the fuel, 
and if the temperatures of CO and air be sufficiently high, 
it will burn to CO2. Part of the air for what may be called 
the secondary combustion will always work its way through 
the fuel bed, because it is impossible to bring all oxygen in 
the air passing through into contact with carbon of the fuel. 
The remainder of the air required is generally admitted 
through the fire door and is heated by passing over the front 
part of the fuel bed. If too great a quantity of air is ad- 
mitted in either way its temperature may be so low as to 
cool the CO below its temperature of ignition and thus fail 
to accomplish the object sought. 

It has been shown that the combustion of pure carbon 
with the theoretical air supply would give gases containing 
about 21 per cent by volume of CO2. If the combustible of 
real coal be assumed to consist entirely of carbon, the same 
proportion of CO2 would result from ideal combustion. 
Practically, it is so difficult to bring the oxygen of the air 
into contact with the carbon of the fuel that a large excess 
is always used, the excess coefficient ranging from about 
1.3 to over 2 and averaging about 1.5 to 1.7 under very 
good conditions. The latter figures correspond to percent- 
ages of CO2 of 14 and 12 respectively, but a value as low 
as 10, which corresponds roughly to an excess coefficient 
of about 2, is not at all uncommon and is generally regarded 
as a very good result except in the largest plants. 



310 STEAM POWER 

(b) Complete Combustion of Hydrocarbons. The hydro- 
carbons which appear as volatile matter in the proximate 
analysis are practically all distilled from the fuel, as it is 
heated in the furnace before ignition in the same way as 
when making a proximate analysis. If they are to be 
completely burned they must be mixed with the requisite 
quantity of air after distillation and both the vapors and the 
air must be maintained at a sufficiently high temperature 
until combustion is complete. Part of the air for the com- 
bustion of distilled volatile niters through the fuel bed 
and the rest must be admitted through the fire door or in 
some equivalent manner. 

If the flame formed by burning hydrocarbons is allowed 
to come in contact with cold surfaces, as, for instance, 
the heating surfaces of the boiler, the gases are cooled 
below the temperature of ignition and combustion ceases. 
This results in the deposit of soot (unburned carbon) upon 
the heating surfaces of the boiler and in the carrying of 
soot and unburned hydrocarbons up the stack. The soot 
and some of these hydrocarbons form the unsightly smoke 
so familiarly associated with some stacks. 

Or, if the air supply is at a sufficiently high temperature, 
but is insufficient in quantity, the hydrocarbons are in- 
completely burned and smoke results. 

The formation of smoke can be conveniently studied 
by means of the ordinary kerosene lamp. Such a lamp 
operates by burning hydrocarbons of the same general 
character as those distilled from solid fuels. The hydro- 
carbons are drawn up by the wick in the form of liquids, 
are vaporized by heat near the top of the wick and then 
combine with oxygen from the atmosphere to give the 
luminous kerosene flame. 

If the flow of kerosene and the air supply are properly 
adjusted and if the temperature is high enough, the com- 
bustion results in the formation of invisible and practically 
odorless gases. If, however, the air supply be decreased 



STEAM BOILERS 311 

or be greatly cooled, a very smoky and very odorous combus- 
tion ensues. The same result could be obtained by the use 
of too great a quantity of air, a condition often attained 
when the supply of kerosene in the bowl of the lamp is almost 
exhausted. 

The effect of a cold surface is easily seen by inserting 
a cold metallic or porcelain surface into the tip of the flame 
and then withdrawing it. It will be found covered with 
soot. 

(c) Advantages and Disadvantages of Excess Air. It 
has been shown that excess air is practically necessary in 
the real furnace in order to insure against a deficiency at 
any point, and it is thus advantageous in that it makes the 
combustion more nearly complete than would otherwise 
be the case. On the other hand, excess air represents just 
so much excess material to be heated at the expense of heat 
liberated by combustion and hence decreases the maximum 
temperature attained. A sufficiently great supply of excess 
air could so reduce the temperature that even if combus- 
tion were complete very little heat would be made available 
for absorption by the boiling vessel, because the temperature 
attained by the products of combustion would be too low. 

Excess air in large quantities may also result in cooling 
unburned gases before combustion to such an extent as to 
make the completion of combustion impossible. 

141. Hand Firing. The commonest type of furnace is 
that shown in Figs. 181 and 182, and the commonest method 
of hand firing consists in spreading a layer of fuel as evenly 
as possible over the entire surface of the fuel bed as often 
as required to replace the fuel burned away. At such inter- 
vals as experience shows to be necessary the fire is cleaned, 
that is, the ashes are worked out from under the fuel by 
means of slice bars, so that practically nothing but live 
fuel resting on a thin layer of ash remains behind. 

This method is open to many serious objections; the more 
important are: 



312 STEAM POWEE 

1. There is a gradual increase in thickness of fuel bed 
from the time of one cleaning until the time of the next. 
This gives a constantly changing set of requirements for 
the proper proportions of air entering below and above 
the fuel bed and a constantly changing resistance to flow 
of air through the bed, so that great skill is necessary if the 
best conditions are to be maintained throughout 

2. There is always a tendency for a fuel bed to burn 
faster at some points than at others, due to the accidental 
distribution of fuel, ash and air. Where " holes " are 
formed in this way large quantities of comparatively cold 
air can pass through with the consequences already enumer- 
ated. It takes considerable skill and watchfulness on the 
part of the fireman to prevent the formation and continued 
existence of such holes. 

3. The firing door must be opened wide every time 
that fuel is to be fired, that is, at intervals varying from two 
or three minutes to fifteen or more, depending on load, 
character of fuel, etc. While the door is open large quanti- 
ties of cold air readily flow into the furnace and cool down 
all parts of it, and a proportionately smaller amount will 
ordinarily pass through the fuel bed. The result of this on 
the flue gases and operation of the boilers has already been 
considered, but there is another result of equal or greater 
importance. As a consequence of this action the volatile 
hydrocarbons distilled off from the freshly fired fuel, which 
are themselves at a comparatively low temperature, are 
surrounded on all sides by cooled walls and come in contact 
with cold air only. The chances of their burning completely 
are very slight, and a great part of these volatilized materials 
passes off unburned as invisible gas and as smoke. Ob- 
viously the greater the volatile content the greater the dif- 
ficulty, so that anthracite causes least trouble in this way, 
while most bituminous coals give heavy black smoke when 
burned under these conditions. 

The cooling down of the interior of the furnace during 



STEAM BOILERS 313 

firing is accompanied by the covering of the fuel bed with 
cold fuel, so that, for the time being, very little radiant 
heat enters the boiling vessel, and the gases which come in 
contact with its surface are comparatively cool. The 
maintenance of a constant steam pressure under these con- 
ditions is practically impossible, but the difficulties can be 
partly overcome by very frequent firing of small quantities, 
so that the door is open a very short time and also that the 
layer of fuel is very thin and does not cut off much heat. 

4. The cleaning of the fire necessitates keeping the 
fire door open for several minutes, with results of the same 
variety as those just enumerated. 

Summing up these difficulties, they divide themselves 
into two classes— those which can be almost or entirely 
eliminated by skill of a very high order and those which are 
inherent and cannot be eliminated by skill. It will also 
b£ observed that all should give more trouble with fuels 
high in volatile than with those of the anthracite variety, 
both as to incomplete combustion and to the formation 
of smoke. 

Several other methods of hand firing have been proposed, 
particularly for use with bituminous coals, and some of 
them have been successfully utilized in isolated instances. 
Nearly all depend upon covering only part of the fuel bed 
at one time and, by alternating the parts covered in this 
way, fresh fuel on one part of the bed is coked while air is 
heated by coming in contact with the uncovered incandescent 
part of the bed and is therefore in proper condition to burn 
more perfectly the volume of hydrocarbons being distilled 
off. These methods are all good, but they involve a great 
deal of careful work and a high degree of skill on the part of 
the fireman. 

Other methods of eliminating some of the difficulties 
depend upon modifications of the furnace and air supply. 
Most attempt to entirely surround the fuel and the gases 
given off with heavy masses of brick work and tile, so that 



314 



STEAM POWER 



enough heat will be stored during incandescent periods 
to tide over the periods of cooling. Some forms have 
combined with this idea a series of air ducts in the brick 
work so arranged that air on its way to the furnace passes 
through these ducts and is heated. In some cases the air 
supply is automatically controlled and more air is supplied 
above the fire during the period of distillation, or coking, 
as it is called, than during the following period, when the 
coked coal is brightly incandescent and little volatile 
matter is present. 




Fig. 183. 



In some hand-fired furnaces which are intended for use 
with bituminous coals that give a long flame the parts of the 
boiling vessel within range of the flames are covered with 
tiles. This prevents impingement of unburned gases upon 
cool surfaces and thus tends to prevent the formation of 
smoke and incomplete combustion. 

Carrying this principle to its logical conclusion results 
in the installation of the grate in a firebrick chamber in 
front of the boiler-setting proper, as shown in Fig. 183. 
Such a device is known as a Dutch oven and is often very 
efficient in totally or partially preventing the formation of 



STEAM BOILERS 315 

smoke. It does not, however, give as high an economy 
as might be expected, because a great part of the radiant 
heat of the fire does not reach the boiler surfaces and because 
the large external surface results in great radiation losses 
to atmosphere. 

Another interesting modification consists of reversing 
the direction of the draft, that is, the direction in which the 
air passes through the fuel bed. The type of furnace al- 
ready described is known as an updraft furnace, because the 
air passes upward in flowing through the bed. The modi- 
fied type here referred to is called a downdraft furnace, 
because the air flows downward in passing through the fuel. 

In downdraft furnaces the coal is fired on top of the 
grate as in other types, but the air is admitted above, flows 
downward toward what would normally be the ashpit, 
and from there on over the heating surfaces of the boiler. 
Fresh coal fired on top of the incandescent bed in such a 
furnace distills as in other types, but the volatiles are mixed 
with the entering air and are carried downward through the 
hot bed so that ideal conditions for combustion are more 
nearly attained. In some forms there is a second updraft 
grate beneath the downdraft grate. This second grate 
receives partly burned coals falling through from the upper 
grate and holds them until combustion is practically com- 
pleted. 

In downdraft furnaces the grate bars are generally made 
of pipes, and water, from the boiler or on its way to the 
boiler, is circulated through them. If this were not done 
the grates would quickly warp out of shape and ultimately 
burn away because of the high temperatures to which they 
are subjected. 

142. Mechanical Grates. In order to overcome the dif- 
ficulties arising from opening the doors for the purposes of 
cleaning the fire, numerous so-called rocking, shaking, 
self -cleaning, or dumping grates have been developed. 
These are generally built up of grate bars which have a 



16 



STEAM POWER 



rough T or an inverted L section with the upper horizontal 
branch of the T or inverted L slightly rounded, as shown in 
Fig. 184. These bars are arranged in groups with their 
longitudinal axes running across the grate, and they are so 
supported that they can be rocked about a point in the verti- 
cal leg of the T or L by means of levers located at the front 
of the boiler. By rocking the bars the lower part of the fuel 
bed which has been burned to ash can be dropped into the 
ash pit, while the upper part is sufficiently agitated to close 
up holes which may have formed, and this can all be done 




/^—-^-^ 



Fig. 184. 



with the doors closed. Or, if desired, part or all of the fuel 
bed can be dropped into the ash pit by a similar rocking 
motion. 

143. Smoke and Its Prevention. An idea of the reasons 
for the formation of smoke will have been obtained from the 
preceding paragraphs. A reasonably skillful fireman should 
have little difficulty in burning anthracite coals in the simpler 
forms of furnaces without smoke, but it is almost impossible 
to commercially burn many of the varieties of bituminous 
coals in this way without the formation of excessive volumes 
of dense black smoke at intervals immediately following 
each firing. 



STEAM BOILERS 317 

Aside from all aesthetic and sanitary considerations, 
smoke is undesirable because it represents poor furnace 
conditions and waste. The actual loss of carbon in visible 
smoke is generally almost negligible in comparison with 
the other losses in the form of unburned hydrocarbons, 
the lowered initial temperature, etc. All of these losses 
combined represent a waste of considerable magnitude. 

The proper method of smoke elimination is not the 
combustion or removal of smoke already formed, but it 
is the burning of fuels in such ways as not to form any 
appreciable quantity in the first place. To accomplish 
this end the following must be achieved : 

1. Coal must be fired continuously and uniformly 
without the opening of doors which admit cold air to the 
furnace. 

2. Volatiles must be distilled continuously and uni- 
formly and in such a place that they are given ample oppor- 
tunity to mix with proper proportions of air and to burn 
completely before coming in contact with cool surfaces. 

3. The air supply must be properly controlled and 
tempered to meet the demands of the fuel both in and 
above the bed. 

4. The fire bed must be worked continuously and 
uniformly so as to eliminate ashes as rapidly as formed 
and to maintain a bed of uniform depth and condition. 

Some of these necessary conditions can be attained by 
the use of the various forms of hand-fired furnaces already 
described but, even in the hands of skillful and industrious 
men, it is impossible to meet all of them. Mechanical 
stokers which more nearly approach the ideals set have 
therefore been developed and are widely used. 

144. Mechanical Stokers. These mechanical devices are 
useful for two reasons — they eliminate a great deal of labor 
and they make possible the burning of many varieties of 
refractory fuels without the formation of excessive quanti- 
ties of smoke. 



318 STEAM POWER 

Despite the good results which can be achieved by 
their use, mechanical stokers are not installed in small 
plants as often as might be expected. This is because 
good stokers are very expensive in comparison with hand- 
fired furnaces and, despite economy of fuel, do not generally 
show a financial saving unless their use eliminates the 
services of several firemen. 

It is generally assumed that one man can care for water, 
coal and ashes for about 200 boiler horse-power or can 
handle coal only for about 500 boiler horse-power. Expe- 
rience has shown that one man can care for about 2000 to 
5000 boiler horse-power when the boilers are equipped with 
good stokers and coal-handling apparatus. 

Financial calculations will generally show stokers to 
be profitable investments for plants of 2000 or more boiler 
horse-power. Where they are installed in smaller plants, 
the absolute necessity of eliminating smoke or the use of 
very poor varieties of coal have generally dictated their use. 

Mechanical stokers can be roughly divided into two 
types, those which duplicate hand spreading of fuel and 
are known as sprinkler stokers, and those which supply 
fuel at one or more points and work it progressively toward 
the ash end of the apparatus as it burns. The first type 
has not been widely installed, though it is possible that 
it may meet with more popular approval after further 
development. 

Stokers of the second type may be roughly divided 
into five classes, which are 

1. Chain grates. 

2. Inclined stokers. 

3. Underfeed stokers. 

4. Combinations of above. 

5. Powdered coal stokers. 

A chain grate, as made by the Illinois Stoker Company, 
is illustrated in Figs. 185, 186, 187, and 188. It consists 



STEAM BOILERS 



319 




O 



320 



STEAM POWER 



of a broad chain made up of a great number of small links 
and carried on toothed wheels and roller wheels supported 
in a frame which can be wheeled into position within the 




Fig. 186. — Sprocket and Links of Illinois Chain Gnte. 





TOP VIEW OF CHAIN 

SHOWING DISTRIBUTION OF AIR oPACES 



BOTTOM VIEW OF CHAIN 

SHOWING ROLLERS FOR DRIVING-SPROCKET 
ENGAGEMENT 



Fig. 1S7. 



boiler setting. The general arrangement of the chain and 
rollers is shown in Fig. 185; details of the front or driving 
rollers and of the links are shown in Fig. 186; a top and bot- 
tom view of part of the chain is given in Fig. 187; and 
Fig. 188 is a perspective view of the frame showing the 



STEAM BOILERS 



321 



tracks on which it may be rolled into and out of the boiler 
setting. 

The chain is driven slowly in the direction indicated 
by the arrows in Fig. 185 by power applied, through worm 
gearing, to the shaft of the toothed wheels at the front 
of the stoker. Coal feeds automatically from the hopper 
by gravity and is carried into the combustion space by 
the moving chain, the thickness of the bed being controlled 




Fig. 188.— Framework of Illinois Chain Grate. 



by the height of the adjustable gate shown. As the fuel 
enters the furnace it passes under the coking arch, which 
spans the entire front part of the grate and which is main- 
tained at a high temperature by heat radiated from the in- 
candescent fuel nearer the inner end of the grate. The 
volatiles are distilled from the fresh coal by heat received 
from this arch and are heated and mixed with air at this 
point. The coked fuel is then carried on into the furnace 
and burned, the refuse being discharged at the bridge wall. 
If the thickness of bed and speed of chain travel are 
properly adjusted, all of the fuel can be coked before pass- 
ing out from under the arch and can be burned almost 



322 



STEAM POWER 



completely before reaching the bridge wall, so that prac- 
tically ashes only will be discharged. 




The apron shown at A in Fig. 185 is used to prevent 
the free passage of air to the part of the chain carrying 



STEAM BOILERS 



323 



practically nothing but ash, as this would result in excessive 
dilution of the products of combustion. 

A stoker of this type installed under a horizontal return- 
tubular boiler is shown in Fig. 189. In the illustration part 
of the side frame of the stoker is broken away m order to 
show the chain and its roller guides. The eccentric shown 




Agitator 
Lock-Nut 



Connecting-Hod 



FlG 190 .— Details of Feed Mechanism, Roney Stoker. 

near the top of the front of the boiler drives the chain 
through an arm of adjustable length, which makes possible 
the control of the speed of chain travel. 

An inclined stoker with front feed and a step grate, 
known as the Roney stoker, is shown in Figs. 190 and 191. 
The fuel is fed out of the hopper and onto the dead plate 
by means of the reciprocating pusher. From the dead 
plate it is pushed down upon the grate bars by the follow- 
ing fuel These bars are rocked mechanically so that their 



324 



STEAM POWER 




OQ 



o 

o 



STEAM BOILERS 



325 



tops alternately assume horizontal and inclined positions, 
and this action feeds the fuel downward until it is dis- 
charged onto the dumping grate. The material collect- 
ing on this grate is periodically dropped by hand into the 
ashpit. 

The fuel is coked while passing under the coking arch 




Fig.' 192. — Transverse Section of the Murphy Stoker. 



and the coked material is practically completely burned 
by .the time it has traveled down the grate. The volatiles 
are mixed under the coking arch with heated air which has 
passed through the grate and with heated air forced in 
above the fuel by steam jets. 

An inclined stoker of the side-feed type with bar grates, 
known as the Murphy stoker, is illustrated in Figs. 192, 
193 and 194. This stoker is provided with two coal- 



326 



STEAM POWER 



magazines or hoppers which are placed horizontally in 
the side walls of the boiler setting and feed fuel onto the 
inclined grate bars, Fig. 192, which carry it downward 
toward the lower point of the V formed by the grates. 
The grate bars, Fig. 194, are alternately fixed and mov- 
able, the movable bars being hung from above and their 




Fig. 193. — Longitudinal Section, Murphy Stoker. 

lower ends being moved up and down by power furnished 
by a small steam engine. 

A toothed bar arranged for rotation by hand or by power 
is located at the bottom of the V and is used for grinding 
up ash and clinker which is too large to fall through into 
the ash pit. This bar is kept cool by making it hollow 



STEAM BOILERS 



327 



and connecting one end to the smoke flues or stack so 
that air is constantly drawn through it. 

The location of the coking arch and the method used 
for supplying warm air should be evident from the figures. 

A stoker of this variety is shown in place under a hori- 
zontal water-tube boiler in Fig. 195. 

An underfeed stoker made by the American Stoker Com- 



Slationary 
Grate Bar 




Movable 
Grate Bar 



Fig. 194.— Grate Bars of Murphy Stoker. 



pany is shown in Figs. 1C6 and 107. Coal is fed from the 
hopper onto the reciprocating bottom B by means of the 
reciprocating pusher P. The latter forms the bottom of 
a trough as shown in Fig. 197, and its reciprocating motion 
feeds the coal upward and out of this trough so that it 
spills over onto the inclined grate bars. The reciprocating 
motions are all obtained from the direct-acting steam cylin- 
der shown. 

The inclined grate bars are alternately fixed and mov- 



328 



STEAM POWER 



able, the movable bars sliding back and forth at right angles 
to the trough under the action of horizontal rocking 




— 
ft 



ft 
H 
3 



bars R. This action gradually feeds the fuel downward 
and toward the side of the furnace, the refuse finally land- 
ing on the dumping trays shown. 



STEAM POWER 



329 




Fig. 196.— Longitudinal Section of Class E American Stoker. 




Fig. 197.— Cross Section of Class E, American Stoker. 



330 



STEAM BOILERS 




STEAM BOILERS 



331 



Air enters the duct below the trough through the 
adjustable gate G, controlled by crank C, and part of it 
passes out through holes H near the top of the trough 
Fig. 197. The remainder passes down through the hollow 




=i 



Fig. 199. — Taylor Stoker Under Horizontal Water-tube Boiler, 



grate bars and into the heated air box from which it flows 
upward between the grate bars. 

It will be observed that the coal is fed onto the grate 
from below, so that all volatiles distilled off must pass up- 
ward through the incandescent fuel before entering the 
space above the fuel bed. Part of the air which is to burn 



332 STEAM POWER 

this volatile matter also passes through the fuel bed and 
the remainder flows over the incandescent fuel from the 
opening shown near the hopper in Fig. 196. The air and 
volatiles are thus raised to a high temperature and well 
mixed, and the operation is continuous and uniform, all 
tending to facilitate smokeless combustion. 

Another variety of underfeed stoker known as the 
Taylor stoker is shown in Fig. 198 (a), (b) and (c). 
This stoker is built up of alternated retorts and air 
boxes, the proper number to give the desired width 
of stoker being used. Coal is fed from the hoppers 
into the retorts by the upper ram or plunger shown 
in Fig. 198 (6) and part of it is again pushed forward by 
the lower ram or plunger. The stroke of the lower 
plunger can be regulated and in this way the relative 
quantities of coal pushed forward in the upper and lower 
parts of the retorts can be controlled. The coal spreads 
over the tuyere blocks which form the inclined tops of 
the air boxes and forms a comparatively even, inclined 
layer of fuel. 

Coking proceeds under the incandescent fuel which 
forms the upper surface of this layer, and the volatiles 
mix with air entering through the hot tuyeres and pass 
upward through the hot fuel above. 

In this stoker advantage is often taken of the fact that 
the draft (pressure of air) required with underfeed stokers 
is so great that it can be more economically attained by the 
use of a fan than by the use of a stack. The fan and the 
coal-feeding plungers are both connected to one engine 
and the speed of this engine is automatically controlled by 
the steam pressure within the boiler. As this pressure 
decreases the engine speeds up, thus delivering more coal 
and air and as the pressure increases the engine slows down 
with opposite results. By properly fixing the travel of the 
plungers initially, the best relative proportions of air and 
coal are set for the entire range of loads to be carried and the 



STEAM BOILERS 333 

variation of both is thereafter in approximately the same 
proportions. 

A stoker of this type in position under a horizontal 
water-tube boiler is shown in Fig. 199. A double-ended 
arrangement of Taylor stokers as used under very large 
water-tube boilers is shown in Fig. 200. 

Powdered-coal stokers have been invented in great 
number and are successfully used in several of the indus- 
tries. They have not, however, been extensively used 
under steam boilers, although isolated installations have 
been reported as giving satisfactory results. 

In all cases, bituminous coal is crushed to a fine powder 
and injected into the furnace with the necessary air for com- 
bustion, the air under pressure generally being made to 
mechanically entrain the coal dust and carry it into the fur- 
nace. The mixture of fine coal and air gives an intensely 
hot blow-pipe type of flame, and firebrick and tile are 
generally used to prevent it from impinging directly upon 
metallic parts. 

Oil firing is essentially a mechanical, rather than a manual 
process, and while oil burners are not ordinarily under- 
stood as belonging to the class of mechanical stokers, they 
have all the- essential characteristics of such apparatus. 

To burn oil successfully under a boiler it must be finely 
atomized and mixed with the necessary quantity of air, 
and there must be sufficient open space within the furnace 
for the free development of the flame and the completion 
of combustion before impingement on cool surfaces. 

Oil-burning furnaces are generally given a rather large 
volume; considerable firebrick is used in such ways as to 
give incandescent walls and baffles to assist ignition and 
combustion, and all heating surfaces are arranged so that 
they are not in the direct path of the flame. 

The atomization of the oil is effected in two distinctly 
different ways. In some forms of burners it is brought 
about by mechanical means, the oil being pumped through 



334 



STEAM POWER 




Fig. 200.— Double-ended Arrangement of Taylor Stoker under Sterling 
Type W. Boiler. 



STEAM BOILERS 335 

a nozzle of some sort which is so shaped that the issuing 
jet breaks up into a great number of very small particles. 
In other forms, steam is used to break up the jet, the steam 
and oil entering the body of the burner separately and later 
coming into contact in such a way that the oil is literally 
torn apart by the steam. This form of burner has been 
more extensively used in the United States than has the 
former. 

Oil burning shares with the burning of powdered coal, 
the property of permitting very accurate regulation of the 
air supply to suit the quantity of fuel being burned. The 
excess coefficient may therefore be maintained at a low value 
and the initial temperature may be made correspondingly 
high. Part of the advantage thus gained over the com- 
moner methods of coal firing is, however, counterbalanced 
by the quantity of steam used for heating and pumping the 
oil and for atomizing in some forms of burners. 

Both oil burning and powdered-coal burning can be 
easily made to give smokeless combustion in properly 
designed furnaces and both yield readily to forcing. That is, 
the temporary consumption of excessive quantities of fuel 
to tide over short 'demands for excessive amounts of steam 
is comparatively easily effected. 

145. Rate of Combustion. The rate at which coal is 
burned in a given furnace or on a certain grate is generally 
given in terms of pounds of coal fired per square foot of grate 
surface per hour and is referred to as the rate of combustion. 

The rate at which coal can be consumed is largely 
dependent on the intensity of draft available, that is, on 
the air pressure available for driving air through and over 
the bed of fuel. The higher the pressure available, the 
greater will be the quantity of air which can be supplied 
and the greater will be the quantity of coal that can be 
burned. If it were not for the cost of creating the draft, 
the only limit to increasing the rate of combustion would 
occur when the velocity of the air became so great that the 



336 STEAM POWER 

fuel would be picked up from the grate and carried 
onward into the flues in a partly burned condition. Com- 
mercial drafts give pressure differences above and below 
the fuel bed which range from about 0.1 inch of water to 
as high as 8 ins. In stationary plants the pressures generally 
range from 0.1 to about 0.5 in cases where hand firing is 
employed, and are carried as high as 5 or more inches 
of water with some forms of mechanical stokers. 

The best rate of combustion varies with the type and 
size of fuel, the type and size of furnace, the type and size 
of boiler, the draft and many other considerations. In ordi- 
nary power-plant practice the rates of combustion com- 
mercially used generally fall within the following limits: 
with anthracite, 15 to 20 lbs. per square foot per hour; with 
semi-bituminous, 18 to 22 lbs.; and with bituminous, 24 to 
32 lbs. There is a rapidly growing tendency to exceed these 
values, particularly in the case of large plants. 

As practically all of the volatile is consumed above the 
grate, the fixed carbon content is practically the determining 
factor, since it is this constituent that is burned on the 
grate. This explains the high rate possible with fuels with 
high volatile content. The most economical results are 
generally obtained when from 12 to 16 lbs. of fixed carbon 
are consumed per square foot of grate per hour. 

The figures given above do not represent limiting con- 
ditions. In torpedo-boat practice, where high-draft pres- 
sures are used (from 4 to 8 ins. of water), rates of from 50 
to 120 lbs. are attained. On locomotives, which also use 
high-draft pressures, rates of combustion greatly in excess 
of stationary practice are generally used. 

The capacity of a given boiler, that is, its ability to 
generate steam, increases as the rate of combustion is 
increased, since more heat is thus made available. The 
economy of the combination, that is, pounds of steam 
generated per pound of coal fired, increases until some best 
rate of combustion for the fuel in question is reached, and 



STEAM BOILERS 



337 



thereafter decreases. The variation of economy is, however, 
not very great for a comparatively wide range of combustion 
on either side of the best rate. ' 

Curves giving approximate draft pressures required fcv 
different rates of combustion when different kinds and sizes 
of fuel are hand fired are given in Fig. 201. The sizes 
referred to are explained in Tables XIII and XIV. Table 
XIII also shows the relative increase of ash content as the 






1% 

S3 a 

















/ 












cv/ 




/ 


/ 














vf 






/ 






'/ 






if 


•-5-V 


V 


7 




$pS 


& 








j?f 




V 


$& 




^ 










\ 








ie 












BvU 


&^ % ' Y 









6 10 15 20 25 30 35 40 

Pounds of Coal per Sq. Ft. of Grate Surface per Hour 



60 



Fig. 201. — Draft Required for Different Rates of Combustion with 
Different Sizes and Kinds of Fuel. 



size decreases, there being a tendency toward the concen- 
tration of the ash in the smaller sizes. 

146. Strength and Safety of Boiler. Attention has 
already been called to the fact that the boiling vessel has to 
be designed with two different requirements in view: it 
must be mechanically strong to resist internal pressure and 
it must transmit the maximum amount of heat to the con- 
tained water. 

Spherical and cylindrical surfaces with the pressure act- 



338 



STEAM POWER 



TABLE XIII 

Sizes of Anthracite Coal 

(Sizes larger than pea coal generally too costly for power-plant use.) 





Through Screen 


Over Screen 


Ash Content 


Name. 


with Mesh. 


with Mesh. 


(Average). 




(Inclusive.) 


(Inclusive.) 




Run of mine 


unscreened 


unscreened 




Broken 




2f 

2 




Egg 


2f 


6 


Stove 


2 


11 


10 


Chestnut 


u 


3 

4 


13 


Pea 


3 

4 


1 
2 


15 


Buckwheat No. 1 


1 
2 


4 


17 


Buckwheat No. 2 or rice. . . 


1 
4 


8 


18 



TABLE XIV 

Sizes of Bituminous Coal^ 

(Considerable variation in commercial practice in naming and sizing.) 



Name. 


Through Bars 

Spaced Apart. 

(Inches.) 


Over Bars Spaced 

Apart. 

(Inches.) 


Lump , 




1? 


Nut 

Slack 


11 

3 
4 


4 



ing on the inside of the curve are best adapted to resist 
such pressures, as they already have the shape which the 
pressure would tend to give them. Boilers are, therefore, 
constructed as far as possible of vessels having only spherical 
and cylindrical surfaces. 

Flat surfaces which are poorly adapted to resist such 
pressures as act within a boiler must often be used despite 
their weakness. When incorporated in a boiler they are 
invariably " stayed," that is, braced by being fastened to 
other surfaces by stay bolts and other forms of fastenings. 
Examples will be given later. 

Most of the early designs of boilers and many of the 



STEAM BOILERS 



339 



modern types consist of large cylindrical vessels made by 
riveting together properly shaped steel plates. These shells 
are often traversed from end to end by flues or tubes for 
carrying hot gases and generally have flat ends more or 
less perfectly braced by these tubes and by long tie rods 





Fig. 202.— Lap Joint. 



Fig. 203.— Butt Strap Joint. 



and other braces. Such boilers when in operation are almost 
entirely filled with water and often hold many tons. 

Boilers of these types have been responsible for many 
disastrous boiler explosions, and this fact has led inventors 
to the development of models which should be less dangerous. 
It seems practically impossible to develop a commercial 
boiler which cannot be made to explode to a certain extent 





Fig. 204.— Riveted Plates of Boiler Shell. 



Fig. 205. 



if sufficiently mistreated and mishandled, but much can be 
done to minimize the danger. 

The great weakness of the older forms lies in the riveted 
joints, which can never be made as strong as the plates 
which they fasten together. Two types of joint are in use; 
they are known respectively as the lap joint and the butt 
strap joint. These are shown in Figs. 202, 203 and 204. 



340 STEAM POWER 

So far as a circumferential seam, that is, one running around 
the cylinder as shown in Fig. 204, is concerned, the lap joint 
is perfectly satisfactory and is universally used. With 
longitudinal seams, however, this is not the case. A lap 
joint throws the joined edges out of a true cylindrical 
surface as shown in Fig. 205, and when the vessel is subjected 
to pressure there will be a tendency for the plates to assume 
a cylindrical contour as nearly as possible. This causes 
local bending of the plates on each side of the lines of rivets, 
and the continued repetition of this action ultimately causes 
failure. The conditions are often made still worse by calking 
the joint on a line indicated by a in Fig. 205, that is, by 
hammering the metal at the inner surface of the edge of 
the outer plate into firmer contact with the outer surface 
of the inner plate for the purpose of making a tight joint. 

The butt-strap joint can obviously be made so that the 
joined plates more nearly form a true cylindrical surface. 

Other weaknesses of the older forms lie in the flat surfaces 
used; in constructions which render it possible for sediment 
to collect on heated surfaces and thus permit local over- 
heating of the plate; and, above all, in the very large 
quantity of water contained. 

The disastrous consequences of boiler explosions are 
generally due to the action of the hot water contained within 
the boiler and not to the steam contained at the time rupture 
occurs. The water within the boiler is under steam pressure 
and approximately at steam temperature. Removal of the 
pressure by rupture of the container would enable a great 
part of this water to flash suddenly into steam at the 
expense of its own heat, and this is exactly what occurs in 
the case of a boiler explosion. Local failure causes a sudden 
lowering of pressure, and this results in the formation of 
large volumes of steam which, blowing out through the 
initial fracture, tend to enlarge it, to move the boiler and 
surroundings, and, in general, to do all possible to further 
the rupture and make conditions worse. 



STEAM BOILEKS 341 

From the preceding discussion the requirements for 
maximum safety can be deduced. They are: 

1. The smallest convenient diameter of cylindrical ves- 
sels, so as to decrease the total load on joints for any 
given steam pressure. 

2. The elimination of the greatest possible number of 
riveted joints and the use of butt-strap longitudinal joints 
on all large-diameter, cylindrical vessels. 

3. The substitution of curved surfaces for all flat stayed 
surfaces. 

4. So shaping the boiler that the required extent of 
heating surface may be obtained without enclosing a great 
volume to be filled with hot water when the boiler is 
steaming. 

5. So shaping the boiler that such water as is contained 
therein will be divided up into small masses contained 
within separate vessels connected in such a way that rapid 
flow of all water toward one point of failure is impossible. 

6. So shaping the boiler that no riveted joints shall be 
in the paths of flames and that no sediment can collect on 
metal immediately over flames or exposed to very hot gases. 

7. So shaping the boiler that it shall 
be free to expand and contract with 
changes of temperature, with the least J I 
resultant strain on the different parts. /^PP^ 

These various requirements are most W "f O 

nearly met in the different forms of water- )l ik\ 

tube boilers, some of which will be de- // ][ \ 

scribed in succeeding paragraphs. 

147. Circulatioa in Boilers. If a flask ™ 

of water, such as that shown in Fig. 206, ^jj^~"^^" 
be heated in the manner indicated, the 
water will gradually acquire motion and follow paths 
such as those shown by the arrows in the illustration. The 
heated water will rise in the center of the mass and the 
cooler water will flow downward around the outer surface. 



342 



STEAM Psi WER 



Such motion is called circulation. Rapid circulation within 
a boiler is very desirable, since it brings the maximum quan- 
tity of water in contact with the heating surfaces in a given 
time and hence tends to increase the amount of heat taken 
from those surfaces. It also tends to sweep along an}' 
bubbles of steam or gas formed on such surfaces and to cany 
away any sediment which may have collected, thus pre- 
venting overheating of the surfaces. 

Circulation can be expedited by providing free and 
unrestricted paths for the water so as to guide it in the 
proper directions and by applying the most intense heat 
at the proper point along the path of the water. The tem- 
perature of the water which is subjected to the most intense 



^M&_ 



Er^-3&£ 



Bias- off 



Fig. 207. — Elementary Types of Boilers. 




heat is naturally raised and the water at that point becomes 
less dense than in other parts of the boiler. The formation 
of steam at such points also materially lessens the density. 
As a result of this lowering of density the heated water 
rises and the cooler water descends to take its place. The 
more rapid this exchange can be made, the more steam can 
be generated from a given amount of surface in a given 
time and hence, other things equal, the better the boiler. 

The elements of two common forms of boiler are shown 
in Fig. 207, the arrows indicating the direction of the cir- 
culation and its effect upon the delivery of steam and 
of sediment. 

148. Types of Boilers. In a book of this scope it would 
be impossible to describe all the types of boilers at present 



STEAM BOILERS 



343 



in use. The more important varieties have therefore been 
chosen for description and illustration. 




Two types of internally fired, tubular boilers more 
accurately described as internally fired, upright or vertical, 
fire-tube boilers are shown in Fig. 208. The furnace is 



344 



STEAM POWER 



Tube Sheet - 

Steam Space- 
Exposed Tubes- 
Water Level- 



-Water Column 
and Try Cocks 



Feed Water 
Connection 




Pressure 
Gauge 



-Hand Holes 

(Closed by 
hand hole covers 
wueu in operation) 



Fig. 209.— Large Internally Fired Tubular Boiler. 



STEAM BOILERS 345 

contained within the shell of the boiler and is almost com- 
pletely surrounded with water. The heat radiated from 
the hot fuel is thus almost entirely received by the water 
of the boiler. The hot gases, rising from the fuel bed, 
pass upward through the tubes and, after giving up part 
of their heat to the surrounding metal, enter the smoke 
box and pass directly to the stack. 

Boilers of the type shown in Fig. 208, (a) and (b) are 
called exposed-tube boilers, because the water level is carried 
below the tops of the tubes. The tubes, therefore, extend 
through the steam space and act as imperfect superheaters. 
Boilers of the type shown in Fig. 208, (c) , in which the tubes 
do not enter the steam space, but are entirely covered by 
water, are called submerged-tube boilers. 

Upright tubular boilers of the types shown in Fig. 208 
are built by a number of manufacturers in sizes ranging 
from about 4 boiler horse-power to about 50 boiler horse- 
power. They are self contained, require no setting of any 
kind, and are shipped completely erected. Such boilers 
are very often mounted on trucks or skids and used to 
generate steam for small hoisting and other forms of con- 
tractors' engines. They are also used on steam fire engines. 

The pressure carried in these small tubular boilers is 
generally under 100 lbs. per square inch, but they can be 
built for higher pressures if desired. 

In Fig. 209 is shown a larger type of internally fired 
tubular boiler as made by the Bigelow Company for station- 
ary use. These boilers are similar to those just described, 
but are made only in large sizes, in this case, in sizes ranging 
from 40 boiler horse-power to 200 boiler horse-power. The 
exposed tubes generally give a superheat of about 25° F. 

These large upright boilers can be built to operate with 
a pressure as high as 200 lbs. per square inch and because 
of the small area covered by even the largest sizes, they 
are particularly adapted to locations in which floor space 
is limited. 



346 



STEAM POWEE 



The locomotive type of boiler is shown in Fig. 210. It is 
an internally fired, horizontal, tubular or fire-tube, boiler. 

Such boilers are seldom used for stationary purposes, 
but are universally used on steam locomotives and, in the 
smaller sizes, are often mounted on trucks or skids and used 
for semi-stationary purposes by contractors and others. 
Boilers of this type are built in sizes ranging from 10 boiler 
horse-power or less up to over 100 boiler horse-power for 
general power purposes, while those used on the largest 
locomotives generate over 2000 boiler horse-power. 

The Continental type of boiler, named from the Con- 
tinental Iron Works, is shown in Fig. 211. These boilers 




Handholes 



Fig. 210. — Locomotive Type of Boiler. 



may be described as internally fired, return tubular, with 
semi-external combustion chamber, this chamber being out- 
side of the boiler shell proper but being built as an integral 
part of the boiler and transportable therewith. Boilers of 
this type are built in sizes ranging from about 75 boiler 
horse-power to 300 or more. 

The grates, furnace and ash spaces, and bridge wall are 
all carried within circular, corrugated flues, one flue being 
used in the smaller sizes and two in the larger. The corru- 
gations serve the double purpose of strengthening the flue 
and of exposing added heating surface to fire and hot gases. 

The steam pipe shown just below the steam connection 
at the top of the boiler is commonly used on boilers for the 



STEAM BOILERS 



347 



poipg noipsg dox 




O 



O 



348 



STEAM POWER 



purpose of preventing the escape of excessive quantities of 
moisture with the steam. 

These boilers are very compact in shape and are short 
for their capacity, but they contain a great volume of water. 
They possess the advantages of having a large steam space 
and a very extended liberating surface over which the steam 
separates from the water. 



Uptake 




Tubes 



Fire Doors 



Ash Pits 



Man Hole- 

Fig. 212.— Scotch Marine Type Boiler. 

The Scotch marine type of boiler is shown in Fig. 212. 
It has the same general construction as that just described 
excepting that the combustion chamber is entirely enclosed 
within the water space of the boiler. This chamber is 
built up of flat plates and is held against collapse by numer- 
ous stay bolts. Boilers of this type were until recently 
the standard for marine practice, but they are now being 
replaced in many instances by water-tube boilers of more 
recent design. 



STEAM BOILERS 



349 



Scotcn marine boilers are very economical in the use of 
fuel, are good steamers, and are absolutely self contained. 
They are built in numerous sizes, the smallest having shells 
with diameters of about 6 ft., while the largest diameter 
used is about 16 ft. The largest boilers have three and 
four corrugated furnaces. 

Two types of externally fired, return-tubular (or 
" H.R.T.") boilers are shown in Figs. 213 and 214. The 



Stay Rods- 



Fire 



Tubes 




"Full : 
Front 



Fig. 213.— Horizontal Return-tubular Boiler with " Full Flush Front." 



only essential differences in these two types are in the forms 
of setting and in the methods of suspending the boilers. 
The shell is generally rigidly supported at the furnace end 
and arrangements made to allow for movement of the other 
end with changes of temperature. 

These boilers can be built very cheaply and are therefore 
widely used when their limitations do not prevent. It has 
been found inadvisable to build them in sizes larger than 
200 boiler horse-power or for pressures higher than 150 
lbs. per square inch, and they are generally used in smaller 



350 



STEAM POWER 




STEAM BOILERS 



351 




352 



STEAM POWER 



Fig. 216.— Forged 
Header for Bab- 
cock & Wilcox 
Boiler. 



sizes and with lower pressures. These limitations are set 
by permissible thickness of metal immediately above the 
fire, experience having shown that the 
plates deteriorate rapidly at this point if 
made too thick. 

One form of Babcock & Wilcox water- 
tube boiler is shown in Fig. 215. This 
boiler is built up of sections consisting of 
several tubes joined at the ends by headers, 
and the sections are connected side by side 
at each end to a long horizontal drum. 
The ends of this drum are closed with 
" dished " heads, thus doing away with 
flat surfaces and the necessity for stays 
within the drum. 

A detail of the forged header is shown in 
Fig. 216. It may be regarded as a long 
box of rectangular section with opposite 
walls pierced by circular holes, which has been so distorted 
as to give it a wavy shape. The distortion brings the holes 
into such positions that the tubes when expanded into these 
holes are "staggered," that is, 
do not lie one above the other. 
The general principle in- 
volved in the arrangement of 
these sections or elements and 
the resulting circulation are 
shown in Fig. 217. The location 
of the feed-water inlet and other 
details are shown in F'ig. 218. 
It will be observed that the feed FlG - - Elementary Babcock. 
water enters in such a direction 
and position that it is readily 
picked up by the current of water circulating in the boiler, 
carried toward the rear and down the rear header. During 
this travel it is heated by contact with the hot water in 




& Wilcox Boiler, Showing 
Circulation. 



STEAM BOILERS 



353 



Hand hole opposite end,-] 
of tube, closed by band 
bole cover when in 
operation. 




End of tube 
expanded into 
hole of header. 



Fig. 218. — Details of Babcock & Wilcox Boiler Construction. 



354 



STEAM POWER 



the boiler and most of its impurities are separated out and 
settle in the mud drum at the bottom of the rear header. 




The boiler is suspended by stirrups from beams carried 
by the brickwork as shown in Fig. 215, the tube sections 



STEAM BOILERS 



355 



simply hanging from the drum by the nipples at each end. 
The various parts of the structure are thus free to expand 
and. contract independently as their temperatures change 
and are not bound in any way by the brick setting. 

The steam is collected from a perforated steam pipe near 
the top of the steam space. The baffle shown in Fig. 218 
prevents the steam which rises from the front header from 
carrying the water bodily into the steam space and makes 
the greater part of the water surface in the drum act as 
separating surface. 

The scale which accumulates inside of the tube is removed 
by tools inserted through the hand holes in the front headers 
opposite the ends of the tubes. One of these hand holes 
•and its cover are shown in section in Fig. 218. Soot and 
dust which accumulate on the outer surfaces of the tubes 
are blown off periodically by a steam jet, the necessary 
nozzle and hose being "in- 
serted through the tall and 
narrow side cleaning doors 
shown in Fig. 215 opposite 
<?ach " pass.'' 

A section of the Heine 
water-tube boiler is shown 
in Fig. 219. This boiler con- 
sists of a slightly inclined 
drum with dished heads, two 
sheet-steel headers and nu- 
merous tubes connecting 
these headers. The shape 
of the header is shown in 
Fig. 220, which indicates the 
positions occupied by the tubes and the way in which 
the header is joined to the drum. 

The products of combustion are generally made to pass 
along the tubes by the longitudinal baffles shown, instead 
of across the tubes as in the boiler last described. 




o o o o o o OO o o o 

O o o o o o O o o o o o o oVo O o <>0 

o o o o o o o o o o o 
°A A°A A ° °A°AO ° O ° O ° O ° O 
^o o o oooooooo 
o°o°o°ooo°o° O ° O °< O ° O o o 

<3<DO O O O OJD O O O 

o 6 o 6 o 6 o 6 o 6 o 5 o 6 o 6 o 6 o 6 o 6 o 



„cp o o o o o o 



o o o o o 



Fig. 220.— Front End Elevation, 
Heine Boiler. 



356 STEAM POWER 

The mud drum in this type is located within the boiler 
and consists of a sheet-steel box supported a few inches 
above the bottom of the drum. The feed water enters at 
the front end of this drum and gradually spreads out as it 
is heated by the surrounding water. The greater part of 
the impurities settles to the bottom and is blown off period- 
ically. The warmed water rises and flows out of an opening 
in the top of the box at the front end and there joins the 
circulation of the boiler, traveling toward the rear, down 
the rear header and up the tubes to the front header. 

The interior of the tubes is cleaned of scale through 
hand holes just as in the last boiler. The external surfaces 
are freed of soot and dust by means of a steam jet which 
is introduced through the stay bolts in the headers, these 
bolts being made hollow for this purpose. Since it is not 
necessary to use doors in the side walls for cleaning in this 
type, Heine boilers are often set up in batteries of three or 
more, each interior side wall serving as the side wall of two 
settings. In the case of the boiler last described the neces- 
sity for side cleaning doors makes it impossible to join more 
than two boilers in this way. 

The Heine boiler is supported by standing the front 
and rear headers upon the brickwork of the setting and 
it can therefore expand freely in all directions. 

A section of the Sterling water-tube boiler is shown in 
Fig. 221. This boiler consists of three upper horizontal 
drums connected by short curved tubes and connected to 
a single lower horizontal drum by means of long tubes which 
are curved near the ends. The curves of all tubes are so 
made that the tubes enter the drum surfaces radially, thus 
giving a simple joint which is readily made tight by expand- 
ing the tube into the sheet. 

The feed water is introduced into the upper rear drum, 
and is gradually heated and partly purified as it passes 
downward to the lower drum, in which the greater part of 
the material precipitated from the water is caught and 



STEAM BOILERS 



357 



stored until blown off. From the lower drum the water 
is supposed to pass upward through the front bank of tubes, 
the steam formed passing to the central drum through the 
upper set of short curved tubes, and the water which is 
not evaporated passing to the central drum through the 



Steam Connectic 



Smoke 

CouiK'i tion 



Damper 




Bottom Blow-off 



Fig. 221 — Section of Sterling Boiler. 

lower set of curved tubes. This water passes from the upper 
central drum to the lower and returns through the front 
bank of tubes. Any steam formed in the rear bank of tubes 
or in the rear drum passes to the central drum through the 
short curved tubes connecting the steam spaces. 

The entire boiling vessel is hung from a frame of struc- 
tural steel by means of the upper drums, so that the lower 



358 STEAM POWER 

drum hangs practically free on the tubes. Independent 
expansion of all the members is insured by this method 
of suspension and by the curvature of the tubes, which per- 
mits each one of them to bend to the extent necessary to 
equalize anj^ strains caused by changing temperatures. 

The interiors of the tubes are cleaned by means of tools 
lowered from inside the upper drums and the exterior 
surfaces are blown off by steam jets introduced through 
doors in the brickwork of the setting. 

The Wickes vertical water-tube boiler is shown in 
section in Fig. 222. It consists of an upper and lower cir- 
cular drum, connected by straight tubes expanded into the 
lower and upper heads of the drums respectively. A 
vertical baffle placed in the center of the bank of tubes 
gives an upward path to the products of combustion when 
passing over the front tubes and a downward path when 
passing over the rear tubes. 

The feed water is generally introduced at the rear of 
the upper drum, the circulation being downward in the rear 
tubes and upward in the front tubes. 

The interior surfaces of the tubes are cleaned by tools 
lowered into them by a man standing within the upper 
drum, which is made high enough to make this possible. 
The external surfaces are cleaned by steam jets inserted 
through doors in the brickwork. 

The entire boiler is supported on brackets riveted to 
the lower or mud drum and is free to expand in all directions, 
the brickwork simply enclosing but not confining it. 

149. Boiler Rating. Practically all apparatus which is 
connected with the development of power is given a horse- 
power rating.' In some cases such a method of rating is 
convenient and simple, in others it is inconvenient, irra- 
tional and complicated. The term horse-power, when used 
as a measure of work or power, means very definitely the 
equivalent of 33,000 ft.-lbs. per minute. When, however, 
a certain number of horse-power is used as the rating of a 



STEAM BOILERS 



359 



Steam Connection 




rFeed Water 
Connection 



-Damper 



Connection 



/Bottom 
/Blow^-ofE 



Fig. 222.— Wickes Vertical Boiler. 



360 STEAM POWER 

particular piece of apparatus, it generally means that that 
piece of apparatus, when working at about its best effi- 
ciency, can do what is necessary to make available the 
stated number of horse-power in the plant of which it forms 
a part. 

Thus a boiler rated at a certain horse-power was origi- 
nally supposed to be able to supply the amount of steam 
required by an average engine developing that quantity 
of power and to do this when working at its best efficiency. 
The water rates of engines are, however, so different that 
there is no real connection between boiler horse-power and 
engine horse-power, and it is best to consider the boiler 
horse-power as a perfectly arbitrary unit defined in a certain 
way. 

The American Society of Mechanical Engineers has 
defined the boiler horse-power as the equivalent of the evapora- 
tion of 34.5 lbs. of water per hour from and at 212° F. This 
means the conversion per hour of 34.5 lbs. of water at 212° F. 
into steam at the same temperature and therefore at atmos- 
pheric pressure. 

Each pound of steam generated under these conditions 
requires the expenditure of the latent heat of vaporization 
it atmospheric pressure, which is equal to 970.4 B.t.u. 
according to the latest steam tables. The older tables gave 
965.7. This quantity of heat is known as a Unit of Evapora- 
tion and is abbreviated U.E. The boiler horse-power is, 
therefore, the equivalent of 34.5 U.E. per hour or 34.5 X 970.4 
= 33,479 B.t.u. per hour. 

As practically no power-plant boilers receive their feed 
water at a temperature of 212° F. and convert it into steam 
at the same temperature, it is necessary to convert the 
weight actually evaporated to what it would have been 
from and at 212° F. and then to divide this figure by 34.5 
in order to find the boiler horse-power developed. 

The number of pounds which would have been evaporated 
from and at 212° F. if the same amount of heat had been 



STEAM BOILERS 361 

transmitted is known as the equivalent evaporation, or as 

the equivalent weight of water evaporated into dry steam from 
and at 212° F. 

The method of obtaining the equivalent evaporation 
has been defined by the American Society of Mechanical 
Engineers. The heat given to each pound of dry saturated 
steam produced is to be determined; this is to be multiplied 
by the total weight of dry saturated steam generated per 
hour, and the product is to be divided by the latent heat 
of vaporization at 212° F. Thus, for a boiler receiving 
its feed water at some temperature t f above 32° F., the 
water contains a quantity of heat equal to q f B.t.u. per 
pound, qf being found in the steam table opposite the tem- 
perature tf. Each pound of dry saturated steam leaving 
the boiler carries with it an amount of heat equal to \ for 
the existing temperature. The heat supplied each pound 
in the boiler must therefore be X— q f and, for W pounds 
per hour, the heat supplied would be W(\— q/). The 
equivalent evaporation is then given by 

Equiv. evap. = W ( _ lf . j lbs. per hour. . (101) 

This expression may be regarded as consisting of two 
factors, the weight of dry steam generated per hour, and 
a fraction which will always have the same value for a given 
combination of pressure and feed-water temperature. This 
fraction is called the factor of evaporation, and it is cus- 
tomary to tabulate the various values of the factor of 
evaporation for different common combinations of pressure 
and feed-water temperature. 

It should be noted that the equivalent evaporation as 
defined above gives the boiler no credit for heat given to 
water which leaves the boiler as water, nor does it give 
credit for any superheating. The former may be justified 
by saying that the boiler, as a commercial piece of apparatus, 
is not intended to supply hot water; but many commercial 



362 STEAM POWER 

boilers are expected to supply superheated steam and should 
be given credit for heat used in that waj'. 

Returning now to the boiler horse-power, its value can 
obviously be found for any given boiler by dividing the 
equivalent evaporation per hour by the number 34.5. 

Boilers are supposed to be so rated that they will 
develop their rated horse-power when operating at about 
their best efficiency and will do it with moderate draft 
and reasonably good firing with average fuel. Experience 
has shown that for most boilers the best efficiency is ob- 
tained when an equivalent evaporation of from 3 to 3.5 lbs. 
of water occurs per square foot of heating surface. The 
heating surface is generally taken as the total surface in 
contact with hot gases excepting in the case of tubes. The 
outer surfaces of tubes are generally counted even if they 
be in contact with the water. An equivalent evaporation 
of 3 to 3.5 lbs. per square foot would call for a heating 
surface of from 12 to 10 sq.ft. per boiler horse-power. 

Most water-tube boilers are given 10 sq.ft. of heating 
surface per rated boiler horse-power, and most return- 
tubular boilers are supplied with 11 to 12 sq.ft. Scotch 
marine boilers are generally designed on a basis of about 
8 sq.ft. per rated boiler horse-power. 

The quantity of water which can be evaporated per 
square foot seems to depend to a great extent upon the rate 
at which hot gases can be passed over the heating surface, 
and experiments have shown that from five to eight times 
the ordinary rates of evaporation can be attained if suf- 
ficient fuel can be burned. As the rate of evaporation per 
square foot is increased above the commonly accepted 
value, the efficiency decreases, but the decrease is generally 
small for a considerable increase in rate of evaporation. 
Most pow T er-plant boilers can give from 150 to 200 per cent 
of their normal rating, and some are now being installed to 
operate for long periods at about 200 per cent of what 
would be considered a normal rating. 



STEAM BOILERS 363 

150. Boiler Efficiencies. There are a great many pos- 
sible efficiencies which may be considered in connection 
with boiler tests. The two most commonly used are denned 
by the A.S.M.E., and are: 

1. Efficiency of the boiler 

_ Heat absorbed per pound of combustible burned 
Calorific value of 1 lb. of combustible 

2. Efficiency of boiler and grate. 

_ Heat absorbed per pound fuel 
Calorific value of 1 lb. of fuel* 

The names used are not very well chosen, and it is better to 
call the first the efficiency based on cornbustible and the 
second the efficiency based on coal. The weight of com- 
bustible burned is calculated by subtracting from the coal 
fired the total weight of moisture and the total weight of 
refuse in the ash pit. 

The heat absorbed is by definition the heat absorbed 
by the dry steam made by the boiler, but it seems probable 
that this will also be modified in the near future as suggested 
in a preceding paragraph. 

It is also possible to determine the efficiency of the grate, 
of the furnace, and of the boiling vessel, and this is some- 
times done. 

The best commercial operating values for the efficiency 
of the boiler as a whole, that is, the boiler and grate on the 
basis of total fuel fired, are about 75 per cent for good 
qualities of coal and 80 per cent for oil, but such values 
are generally obtained only in well-equipped plants operat- 
ing on comparatively constant loads. Average commercial 
values generally range from 60 to 70 per cent on a yearly 
basis in well-equipped plants which are carefully operated, 
and many boiler plants are operated at an efficiency of 50 
per cent and less. 

The pounds of water evaporated per pound of coal 



364 



STEAM POWER 



fired generally ranges between 6 and 10, and the equiva- 
lent evaporation per pound of combustible burned will 
generally fall between 8 and 12 pounds. 

151. Effects of Soot and Scale. The flue gases in real 
boilers are seldom clean mixtures of the products of com- 
bustion and nitrogen, as theory would indicate. They 
always contain more or less soot and unburned hydro- 
carbons, as well as some finely powdered ash and fuel. With 
strong draught, very large particles of ash and fuel may be 
carried by the flue gases. 

These materials are partly carried up the stack by the 
gases and partly deposited on the heating surfaces of the 
boiler. Such deposits decrease the conductivity of the 
heating surfaces, and if the deposits are heavy the loss may 
be very great. The results of one investigation on the 
effect of soot are given in Table XV, the values being 
taken from an article published in the Proceedings of the 
Institute of Marine Engineers for the year 1908. These 
values are probably too high, particularly for the thicker 
deposits, but they serve to bring out the fact that a very 
appreciable loss does occur from the presence of such 
deposits. 

TABLE XV 
Effect of Soot Deposits on Boiler Heating Surfaces 



Thickness of Deposit 


Loss of Conductivity 


in Inches. 


in Per cent. 





0.0 


1 

32 


9.5 


1 
16 


26.2 


1 
8 


45.2 


3 
T6 


69.0 



The effect of soot deposits in decreasing the efficiency 
of boilers was used for a long time as a basis for argument 
in favor of certain types of boilers in which the heating 
surfaces were so shaped and located that such deposits 



STEAM BOILERS 365 

formed to a minimum degree and against other types 
less favorably designed from this point of view. Prac- 
tically, however, the removal of such deposits by means of 
steam jets applied at regular intervals is so simple that this 
consideration need be given little weight in the selection 
of a boiler. Provision should always be made, however, 
for the easy use of the jets for cleaning purposes. 

152. Scale. Practically all water available for boiler feed 
contains various salts in solution and it often contains solid 
matter in suspension as well. This material is all deposited 
within the boiler as the water is heated and converted into 
steam. There is thus a gradual collection within the boiler 
of all the solid material brought in by the water. 

In well-designed boilers the greater part of such deposits 
is carried to a part of the boiler in which the metallic sur- 
faces are not exposed to high temperature gases, as, for 
instance, the mud drums in water-tube boilers. It can then 
be drawn off periodically in the form of a thin mud sus- 
pended in water. In practically all boilers, however, some 
of the solid material will be carried to the heating surfaces 
exposed to high temperature gases and deposited there. 
Under the action of heat, the mud-like material gradually 
changes until, in many instances, it forms a very hard, 
stone-like coating on the heating surface. This is known 
as boiler scale. 

Such deposits may cause two kinds of trouble: They 
may decrease the conductivity of the heating surfaces and 
thus decrease the efficiency of the boiler; and, because of 
their location on the water side of the metal, they may per- 
mit the hot gases to overheat that metal, thus weakening 
it. Such overheated metal often " bags " under the high 
internal pressure and may eventually give way with disas- 
trous results. The mechanical structure of the scale seems 
to be the determining factor; scales which are easily pene- 
trated by water have little effect, while those which are very 
dense and non-permeable may cause serious trouble. 



366 STEAM POWER 

Boilers should be blown down periodically to keep them 
as free as possible of scale-forming material, and the}^ should 
be so constructed that scale which has been formed can be 
removed easily. Very efficient tools have been developed 
for removing scale from the interior and exterior surfaces 
of tubes, so that boilers using tubular heating surfaces are 
readily cleaned of scale. 

153. Scale Prevention. Much of the solid material 
carried by water is deposited when the water is heated to 
a temperature of from 150° to 200° F., so that heating feed 
water before it is admitted to the boiler is at least a partial 
preventive in most cases. 

Nearly all of the salts which are soluble in hot water 
and therefore are not deposited when the feed is heated, 
can be made to form insoluble compounds by the addition 
of comparatively cheap chemicals. By the addition of such 
chemicals in the feed-water heaters, or in other apparatus 
specially designed for that purpose, the greater part of the 
solid content of the water can be precipitated before it is 
admitted to the boiler. 

There are a great many " boiler compounds " on the 
market which are intended to be mixed with the water as 
it is fed to the boiler and are supposed to prevent the for- 
mation of scale on the heating surfaces. All they can 
possibly do is to change the chemical composition of the 
solids; they cannot prevent the deposit of these solids within 
the boiler. They are therefore, at best, only an imperfect 
remedy. 

154. Superheaters. Many boiler plants arc now ar- 
ranged to supply steam superheated 25 to 200 degrees Fahr. 
It was shown in an earlier chapter that the use of super- 
heated steam greatly improves the economy of reciprocating 
engines and turbines, and there are also other advantages 
which accrue from its use. 

Superheaters are of two kinds — separately fired and 
built-in superheaters. The separately fired superheaters are 



STEAM BOILERS 367 

enclosed in a brick setting fitted with grate and furnace 
similar to that of an ordinary boiler. The built-in super- 
heaters are installed within the boiler setting so that the 
products of combustion pass over them in flowing through 
the boiler. 

In either type the steam passes through the superheater 
on its way from the boilers to the engines. In the case 
of separately fired superheaters, the temperature of the 
superheated steam is controlled by regulation of the fire 
on the grate of the superheater, but in the built-in type 
regulation in this way is practically impossible, as the fire 
under the boiler must be controlled to suit the demand for 
steam. The control of such superheaters is therefore effected 
either by locating them in such a position that the natural 
variation in the temperature of the gases reaching them 
gives an approximate regulation, or they are installed in 
a separate chamber and hot gases passed over them in such 
proportions as necessary to give the required temperature. 

The Babcock & Wilcox superheater as applied to the 
boiler of "the same make is shown in Fig. 223. The steam 
collected in the dry pipe within the drum passes downward 
to the upper manifold of the superheater and from there 
it flows through the U-shaped tubes into the lower manifold. 
From the lower manifold it flows through the superheater 
stop valve to the engine or turbine. 

The superheater is so located that the hot gases pass 
over it between the first and second passes and there is no 
way of shutting off these gases. Provision, as shown in 
the illustration, is therefore made for flooding the super- 
heater during starting, or when superheated steam is not 
desired. When flooded it becomes heating surface similar 
to that of the tubes below, the steam made passing into 
the drum through the dry pipe. 

The Heine superheater as applied to a Heine boiler is 
shown in Fig. 224. It consists of a sheet-metal header or 
box into which U-shaped tubes are expanded. The steam 



368 



STEAM POWER 



Safely Valve 



Stop Valve 




.Drain 
Valve 



Front 



Gases 
from Furnaces 



Fig. 223. 



-Superheated Steam. from 
Superheater 




Fig. 224.— Heine Superheater. 



STEAM BOILERS 



369 



enters the bottom of the header and is guided by dia- 
phragms in such a way that it passes through the lower set 
of U-tubes, returns to the header, passes through the upper 
set of tubes, and then leaves the superheater at the top. 




Fig. 225.— H.R.T. Boiler and Foster Superheater. 



This apparatus is installed in a brick chamber built into 
the boiler setting and connected with the furnace by a flue 
(not shown) in the brick side wall. A damper controls the 
flow of hot gases to this chamber and the degree of superheat 
is controlled by the position of this damper. 




Headers 



Fig. 226. — Element of Foster Superheater. 

The Foster superheater is shown installed in the setting 
of an H.R.T. boiler in Fig. 225 and the details of the con- 
struction of one element are shown in Fig. 226. The core 
is used to spread the steam in a thin stream, thus bringing 
it into better contact with the heating surface. The fins 



370 STEAM POWEK 

on the exterior of the element are used for the purpose of 
getting a more extended metallic surface in contact with 
the hot gases. 

155. Draft Apparatus. Attention was called in a pre- 
ceding paragraph to the fact that there must be a difference 
of pressure between the spaces below and above the fuel 
bed in order to cause the necessary air to flow through the 
bed. This difference of pressure is called the draft. 

As a matter of fact, a slight difference of pressure is 
required to cause the flow of gases through any part of the 
boiler and the drop in pressure through the fuel bed is only 
part of the total draft required. 

The draft may be created in two distinctly different 
ways. It may be caused by a chimney or stack, and is 
then known as natural draft, or it may be produced by fans 
or blowers, in which case it is called mechanical draft. 

(a) Chimneys or Stacks. Stacks are practically always 
used in small plants because of the simplicity resulting from 
their use and because the interest on the investment com- 
pares favorably with interest on investment plus cost of 
operation for mechanical draft. In large plants fitted with 
some types of mechanical stokers, or where fuel is to be 
burned at a high rate, or where the flue gases are to be used 
for heating feed water, mechanical draft is generally installed. 
A stack of some sort is necessary even though mechanical 
draft be used, because the products of combustion must be 
discharged at a sufficient elevation to prevent their being 
a public nuisance. 

A chimney serves to cany away the hot products of 
combustion and when in operation is filled with a column of 
gases with higher average temperature than that of the 
surrounding air. As a result the density of gases within 
the stack is less than the density of the outer air and the 
gas pressure at the bottom of the structure is less inside 
the stack than it is outside. If an opening is made at this 
point, the external air will therefore flow in. By arranging 



STEAM BOILERS 



371 



the apparatus as shown in Fig. 227, the temperature 
of the air flowing into the bottom of the stack is raised 
as it passes through the furnace and the flow is thus made 
continuous. 

The height of the chimney determines the draft created 
by it with flue gases of a given temperature, and, with any 
given height, the area determines the quantity of gas which 




Atmospheric Pressure 
less amount caused by 
low density of hot gases 
in stack. 



Fig. 227. — Diagrammatic Arrangement of Stack. 



can be carried off in a given time. The proportions of 
chimneys can be determined from rational formulas based 
on theoretical considerations, but it is necessary to assume 
values for a number of constants and a proper choice de- 
pends largely upon experience. 

As a result, all but the more important chimneys are 
generally designed on an empirical basis and many formulas 
have been developed for this purpose. One of the most 
common methods of design is to choose the height in accord- 



372 STEAM; POWER 

ance with the values given in Table XVI, and then to 
determine the sectional area according to an empirical 
assumption or formula. 

TABLE XVI 

Common Heights of Chimneys 

(Applicable! to plants smaller than about 700 H.P. Larger installations should 
have stacks of from 150 to 175 feet in height unless local conditions call for 
greater height.) 



Character of Fuel. 


Height above Grate in Feet. 


Free-burning bituminous 


80 


Anthracite, medium and large sizes 

Slow-burning bituminous 


100 
120 


Anthracite, pea size 

Anthracite, buckwheat sizes 


130 

150 



Thus, some designers simply assume the sectional area 
at the top of the stack equal to about one-ninth of the grate 
area for anthracite coal and equal to about one-seventh of 
the grate area for bituminous coal. Others use a formula 
developed by William Kent, which is based upon the 
assumption that the stack should be large enough to carry 
away all the gases resulting from the combustion of 5 lbs. 
of coal per rated boiler horse-power per hour. This formula 
gives the boiler horse-power which the stack can serve 
and is 

H.P. = 3.33(A-0.6VZ)VIf, . . . (102) 
in which 

H.P. = Rated boiler horse-power; 

A = Internal sectional area in feet of circular or square 

chimney; 
H = Height above grate in feet. 

(b) Mechanical Draft. Fans can be so used as to force 
air into the ash pit, that is, to raise the pressure on the 



STEAM BOILEES 373 

entering side of the fire. In such cases the equipment is 
said to give forced draft. Or fans may be installed at the 
discharge end of the flues and may "draw " the gases 
through the boiler by lowering the pressure within to a value 
below that of the external atmosphere. Such an instal- 
lation is said to give induced draft. 

Forced draft suffers from the disadvantage that the 
pressure within the furnace is greater than atmospheric and 
hot gases may therefore be blown out when the fire 
door is opened. On the other hand, the fan handles only 
cool air instead of hot products of combustion as in 
the case of induced draft and its useful life is therefore 
much longer. Forced draft is much more common than 
induced draft. 

Several arrangements giving balanced draft have been 
developed. With such apparatus a pressure equal to atmos- 
pheric is maintained above the fuel bed and no hot gases 
are blown out through the firing door. 

PROBLEMS 

1. The equivalent evaporation of a boiler during a certain 
test was 3450 lbs. per hour. What boiler horse-power was de- 
veloped? 

2. A water-tube boiler with 5000 sq.ft. of heating surface 
and rated in the ordinary way gave an equivalent evaporation 
of 25,875 lbs. per hour. At what per cent of rating was the 
boiler operating? 

3. A certain boiler produced 3500 lbs. of dry steam in one 
hour from feed water at a temperature of 50° F. The steam 
pressure was 200 lbs. per square inch gauge. What was the 
equivalent evaporation? 

4. A boiler receiving water at a temperature of 250° F. con- 
verts it into superheated steam at a pressure of 210 lbs. per square 
inch gauge and a temperature of 580° F. The boiler produces 
20,000 lbs. of steam per hour. What is the equivalent evaporation 
if the boiler is given credit for all the heat given the material 
passing through it? What boiler horse-power is developed? 

5. A boiler produces 7.5 lbs. of dry steam per pound of coal 
fired. The feed-water temperature is 80° F. and the steam pres- 



374 



STEAM rOWEE 



sure is 125 lbs. per square inch absolute. What is the equivalent 
evaporation per pound of coal? 

6. A boiler is supplied with coal which has a calorific value 
of 13,520 B.t.u. per pound. It produces 8 lbs. of dry saturated 
steam at a pressure of 150 lbs. per square inch gauge per pound 
of coal. The feed-water temperature is 70° F. What is the 
efficiency of the outfit? 



CHAPTER XVIII 
RECOVERY OF WASTE HEAT 

156. Waste Heat in Steam Plant. There are^two great 
heat wastes in the steam plant — the waste in the hot gases 
going up the stack and the waste in exhaust steam. The 
magnitude of the stack loss can best be appreciated by 
determining an approximate value for assumed conditions. 
For this purpose assume the fuel to be pure carbon, the 
excess coefficient 1.5, average atmospheric temperature 60° 
F., average stack temperature 600° F., and no moisture in 
the air. The specific heat of the flue gases may be taken 
as constant and equal to 0.24. 

With an excess coefficient of 1.5, the total weight of 
flue gas per pound of carbon burned would be about 18.4 
lbs. and the heat carried up the stack figured above room 
temperature would be 

Stack loss = 18.4X0.24 (600-60). 

= 2380 B.t.u. per pound of C burned (approx.) 

With a calorific value of 14,600 B.t.u. per pound of carbon 
this loss would be equivalent to a little over 16 per cent 
of the total heat in the fuel. 

It would be more correct to use the temperature of 
the steam in the boiler instead of room temperature, because 
the lowest temperature theoretically attainable by gases 
passing through a boiler would be equal to that of the steam 
and water on the other side of the heating surface. Under 
ordinary conditions of operation, this method of figuring 
would give a theoretically avoidable stack loss equal to 
about 50 per cent of the figure obtained above. 

375 



376 STEAM POWER 

The magnitude of the exhaust loss can be similarly 
approximated. Assume for this purpose an engine receiving 
dry saturated steam at 115 lbs. absolute per square inch and 
exhausting it with a quality of 90 per cent at a pressure of 
15 lbs. absolute per square inch. 

The heat above 32° in the entering steam is 1188.8 B.t.u. 
per pound and the heat exhausted per pound is 1053.7. 
The heat in the exhaust represents therefore about 89 per 
cent of all the heat supplied when calculations are made 
above a temperature of 32° F. If a feed-water tempera- 
ture of 60° be assumed and heat quantities be figured 
above that datum the results are practically the same. 

There are always numerous pieces of auxiliary apparatus 
in steam plants such as boiler-feed pumps, circulating pumps, 
vacuum pumps, etc. These are often steam driven and are 
generally very uneconomical in the use of heat, so that they 
throw away in their exhaust steam large quantities of heat 
originally transferred from fuel to water and steam in the 
boiler. 

157. Utilization of Exhaust for Heating Buildings. It 
often happens that steam-power plants are located within 
or in the neighborhood of buildings requiring artificial heat 
during part of the year. In such cases the exhaust steam 
from main and auxiliary engines can generally be advan- 
tageously used for this purpose. Under particularly favor- 
able circumstances, the weight of steam required by the 
plant may equal approximately that required for heating, 
and the greater part of the exhaust could then be turned 
directly into the heating systenl. 

The engines in plants of this character may be regarded 
as reducing valves for the heating system, receiving steam 
at high pressure and reducing the pressure to the value best 
adapted to the heating system installed. If the com- 
paratively small losses arising from radiation from the 
engine, from friction and from the presence of hot water 
in the exhaust be neglected, all heat received by the engine 



EECOVERY OF WASTE HEAT 377 

and not turned into useful mechanical energy is made use 
of in the heating system. The engine may therefore be 
very uneconomical in the use of steam and still not cause 
a waste of fuel, provided always that the heating system 
can absorb all heat exhausted. 

Since the demands of a heating system vary from day 
to day and since there is generally no demand for heat 
during several months of each year, it follows that a high 
degree of skill is necessary in choosing the character of the 
apparatus installed. A compromise is generally made 
between the cheap and uneconomical engine allowable during 
the coldest months and the more expensive and more 
efficient engine desirable when no heating is to be done. 

There are other cases of somewhat similar character. 
In many industries use can be made of exhaust steam for 
the heating of evaporating pans, dye vats, kilns and other 
apparatus. Steam plants of an uneconomical character may 
be very economical financially in connection with such 
industries if all or nearly all of the heat in the exhaust can 
be utilized industrially. 

158. Feed-water Heating. An examination of the steam 
table will show that the total heat above 32° F. per pound 
of steam varies between 1180 and 1200 B.t.u. for such pres- 
sures as are commonly used in boilers. The average tem- 
perature of water as it occurs on the surface of the earth- 
is probably somewhere in the neighborhood of 60°, so that 
the heat above 32° per pound would roughly average 27 
B.t.u. A boiler receiving water at 60° and converting it 
into steam at any of the ordinary pressures must therefore 
supply over 1100 B.t.u. per pound of water. 

This immediately suggests a use for heat in exhaust 
steam. Steam exhausted into very low vacuums has a 
temperature only 10° to 30° higher than the assumed average 
natural feed temperature, but steam exhausted at atmos- 
pheric pressure has a temperature of 212° F. and could 
therefore impart large quantities of heat to water at 60° F. 



378 STEAM POWER 

Since the boiler must supply over 1100 B.t.u. per pound 
of steam made, raising the feed temperature about 11° or 
12° should effect a saving of about 1 per cent in fuel con- 
sumption. By raising the temperature from 60° to 212° 
there should therefore result a saving of approximately 13 
to 14 per cent. 

Other advantages which would accrue from this pre- 
liminary heating of the feed water would be (1) the deposit, 
outside of the boiler, of a large amount 'of the solid matter 
carried by the water, (2) the use of fewer or smaller boilers, 
and (3) the reduction of the strains which occur in the metal 
of some designs when very cold feed water is used. 

Exhaust steam feed-water heaters are divided into two 
types, open and closed heaters. In open heaters the steam 
and feed water are brought into intimate contact in the 
form of jets, sheets and sprays within a vessel of appropriate 
size and shape. They are often called contact heaters. 
When the exhaust steam comes from reciprocating engines 
it always carries in suspension some of the oil used for 
lubricating the engine cylinders. If allowed to enter the 
heater, this oil would mix with the feed water and eventually 
reach the boilers, where it might cause serious damage by 
depositing upon heating surfaces exposed to the fire or to 
very hot gases. Such heaters are therefore always fitted 
with oil or grease extractors when used with reciprocating 
units. When receiving the exhaust from turbines, oil ex- 
tractors are not necessary, as no lubricant is used within 
the steam spaces of such units. 

Closed heaters consist of tubes or coils enclosed within 
a metal vessel. One medium passes through the tubes 
and the other over their outer surfaces. Such heaters are 
therefore often called non-contact heaters. 

As oil is a poor conductor of heat, the exhaust steam 
from reciprocating units should be passed through an oil 
extractor before entering a closed heater in order that the 
heating surfaces may be used to the best advantage. 



RECOVERY OF WASTE HEAT 379 

Exhaust steam feed-water heaters are often divided into 
primary and secondary heaters. This distinction has nothing 
to do with structure, being based entirely on position and 
temperature. Thus there may be available exhaust steam 
at a pressure below atmospheric, as from condensing main 
units, and exhaust steam at atmospheric pressure from non- 
condensing auxiliaries. The lower pressure steam could be 
used to heat the feed water in a primary heater and the 
higher pressure steam could then raise its temperature still 
further in a second or secondary heater. 

The other great waste, that in the stack gases, can also 
be partly eliminated by using some of it to heat the feed 
water. As the highest steam temperature ordinarily avail- 
able in the exhaust system is about 212° F., and as the 
products of combustion leaving the boilers generally have 
temperatures in the neighborhood of 600° to 700° F., it 
is evident that on a basis of temperature the hot gases 
have a decided advantage as a heating medium. On the 
other hand, the specific heat of the hot gases is low, while 
exhaust- steam can give up all of its latent heat with no 
change in temperature, so that on a basis of heat avail- 
able for transmission to the water, the steam has the 
advantage. 

The waste heat in the flue gases is used for feed-water 
heating in devices known as economizers. These generally 
consist of groups of tubes, joined at their ends by headers 
and standing vertically within sheet-metal flues leading 
from the gas passages of the boilers to the chimney. The 
water to be heated is pumped through the tubes on its way 
to the boilers and the hot gases flow over the tubes on their 
way to the chimney. Mechanically operated scrapers are 
arranged to travel up and. down the tubes at intervals and 
keep their external surfaces free of soot and dust, which 
would seriously reduce their ability to transmit heat. 

The feed water supplied the economizer is generally 
first heated in an exhaust steam heater and arrives at the 



380 STEAM POWER 

economizer with temperatures between about 120° and 
200° F., depending upon the kind of heaters used and upon 
the relative quantities of exhaust steam and feed water. 
The economizer discharges the water to the boiler at tem- 
peratures which generally run from about 210° to 300° F., 
depending upon the amount of preliminary heating, the 
extent of economizer surface and a number of other variables. 
The gases which have passed through an economizer 
often have temperatures as low as 250 to 350° F., which is 
generally too low a value to give good chimney draft. 
Plants making extensive use of economizers are therefore 
generally fitted with some form of mechanical draft. . 

PROBLEMS 

1. Determine the heat lost in the chimney gases per pound 
of coal in a plant operating under the following conditions, and 
express the loss as a percentage of the heat value of the coal. 
The coal has a calorific value of 14,000 B.t.u. per pound; the 
temperature of the gases leaving the boiler is 570° F.; 20 lbs. of 
gas result from each pound of coal burned; the mean value of 
the specific heat of the gases is 0.245; and the temperature of 
the air entering the furnace is 75° F. 

2. Determine the quantity of heat which could be obtained 
from the gases of Prob. 1 by using an economizer to reduce their 
temperature to 250° F. What percentage of the heat value of 
a pound of coal does this saving represent? 

3. The boilers of a certain plant produce 100,000 pounds of 
steam per hour when the plant is operating at full load. The 
steam-driven auxiliaries consume 10% of this steam. Steam is 
generated at a pressure of 175 lbs. per square inch gauge, and 
is superheated 150° F. The main units operate condensing and 
the condensate leaves the condensers at a temperature of 75° F. 
The auxiliaries operate non-condensing and exhaust their steam 
at atmospheric pressure and with a quality of 92%. The coal 
used has a calorific value of 13,850 B.t.u. The boiler efficiency 

(Heat given water and steamX . „ r _, 
— - — : — - — ^—. is /5%. 
Heat in fuel supplied / 

(a) Determine the amount of coal which would have to be 

burned per hour if the steam exhausted from the auxiliaries were 

thrown away and make-up water at a temperature of 50° F. were 



RECOVERY OF WASTE HEAT 381 

used in its place. The condensate from the condensers of the 
main unit is assumed to be returned to the boiler after being 
mixed with the make-up water. 

(b) Determine the amount of coal which would have to be 
burned per hour if the auxiliary exhaust were used to heat the con- 
densate from the main units in an open heater and if the operation 
of the plant were so perfect that no make-up water had to be 
added. 



CHAPTER XIX 
BOILER-FEED PUMPS AND OTHER AUXILIARIES 

159. Boiler-feed Pumps. The pumps used for forcing 
the feed water into boilers may be of reciprocating or 
centrifugal construction and may be driven by reciprocating 
steam cylinders, by small steam turbines or by electric 
motors. 

Steam-driven pumps are very wasteful, often using over 
100 lbs. of steam per horse-power hour. It would therefore 
seem more economical to use motor-driven pumps in electric- 
power stations, as the large power units will generate electric 
power with a consumption of from 10 to 25 lbs. of steam per 
horse-power hour and the motor efficiency will generally be 
over 80 per cent. There is, however, another point which 
must be considered. The exhaust steam from small engines 
operating boiler-feed pumps can be used for heating the 
feed water as described in the last chapter, and thus the poor 
economy of these units is of little significance; practically 
all heat exhausted can be returned to the boiler in the 
boiler feed if desirable. As a result of this considera- 
tion, coupled with others of less importance, nearly all 
boiler-feed pumps and other similar auxiliaries are steam 
driven unless there are so many that there would be 
more exhaust steam than could be absorbed by the feed 
water. 

There is at present a marked tendency toward the use 
of turbine-driven, centrifugal pumps for boiler feeding, in 
place of those driven by reciprocating steam units. The 
turbine type has several advantages, the more important 
being : 

382 



BOILER-FEED PUMPS AND OTHER AUXILIARIES 383 

(1) No oil in exhaust steam, so that latter is well adapted 
to use in all forms of feed- water heaters; 

(2) Higher speed because of continuous flow of water 
and continuous rotation of mechanical parts, thus making 
possible great decrease in size for a given amount of work, 
and 

(3) Better pump characteristics for this sort of work. 

The Duplex Steam Pump. The great majority of re- 
ciprocating steam pumps used for boiler-feed purposes are 
of the duplex pattern, one design of which is shown in Figs. 



Steam Valv.e C&est 




Steam End Water En^ 

Fig. 228.— Duplex Steam Pump. 



228 and 229. Two steam cylinders are arranged side by 
side, their piston rods extending into similarly arranged 
water cylinders and carrying water plungers or pistons as 
shown in Fig. 229. As there is no rotating shaft in a pump 
of this kind, the steam valves cannot be operated by eccen- 
trics as is common with steam engines. For the purpose 
of operating these valves, bell cranks, pivoted near the 
center of length of the pump, are provided. These are 
arranged so that the long arm of one bell crank engages a 
collar on the piston rod of one steam cylinder and the short 
arm operates the valve gear of the other steam cylinder. 
The motion of the valve of one cylinder is therefore derived 



384 



STEAM POWER 



from the piston motion of the other cylinder. The steam 
pistons are practically 180° out of phase, one moving out 
while the other moves in. 

Practically no expansion of the steam is obtained in the 
cylinders of pumps of this type. They operate on the 
rectangular cycle described in an earlier chapter and are 
correspondingly wasteful in their use of steam. 



Slide Valves 




Steam. End WatecEnd 

Fig. 229.— Duplex Steam Pump. 

160. The Steam Injector. On steam locomotives and 
in other portable steam plants, as well as in many small 
stationary plants, a device known as a steam injector is used, 
instead of a pump, for forcing feed water into the boiler. 
A simple form of steam injector is shown semi-diagram- 
matically in Fig. 230. 

Steam from the boiler flows through the steam nozzle 
and expands from boiler pressure to a very low pressure, 
thus acquiring a high velocity at the expense of the heat 



BOILER-FEED PUMPS AND OTHER AUXIIJARIES 385 

energy which it brings from the boiler. At the end of the 
nozzle it mixes with water and imparts to that water some 




of its kinetic energy, so that the mixture moves into the 
small end of the delivery tube with a high velocity. By 
the time it has reached that point, practically all the steam 
has been condensed, and, as the sectional area of the delivery 



386 STEAM POWER 

tube increases, the velocity of the liquid decreases with a 
corresponding increase in pressure according to Bernoulli's 
theorem. In properly designed apparatus, the resultant 
pressure is great enough to force the mixture of water and 
condensed steam into the boiler against boiler pressure. 

The space at the end of the steam nozzle is maintained 
at a low temperature by the feed water flowing through 
it and the pressure of the steam is therefore very low at this 
point, being less than atmospheric in most cases. Atmos- 
pheric pressure is therefore able to force water up the suction 
pipe if the " lift " is not too great, and when once started 
such a device can therefore " raise " its own water as well 
as deliver it against pressure. 

It is interesting to note that the efficiency of this appa- 
ratus is almost 100 per cent on a heat basis. All heat not 
radiated from the apparatus is returned to the boiler in the 
mixture of condensed steam and feed water and, as the 
external surface is very small, very little heat is lost by 
radiation. 

161. Separators. Two kinds of separators are used- in 
steam plants: (a) the oil separators already referred to for 
separating oil from exhaust steam, and (6) steam separators, 
which separate water from steam. 

As it is impossible entirely to prevent radiation from 
steam pipes, it follows that condensation will occur in any 
pipe line which carries saturated steam. Water is also 
formed in the cylinders of reciprocating engines not supplied 
with very highly superheated steam, and much of it is 
generally present in the exhaust of the high and inter- 
mediate cylinders of multiple-expansion engines. 

A small amount of water can be passed through the 
cylinder of a reciprocating engine without mechanical 
damage, but it probably causes a loss of heat by cling- 
ing to the walls and assisting in the heat interchanges 
which always occur. Large quantities of water are apt 
to cause mechanical damage, as water is inelastic, and if 



BOILER-FEED TUMPS AND OTHER AUXILIARIES 387 



-Jacket- ofTnsulating 
Material to Decrease 
Radiation Loss. 




Fig. 231. — Steam Separator 



388 STEAM POWER 

more of it is trapped in a cylinder end than can be con- 
tained in the clearance, something must give way when 
the piston reaches the end of its stroke. 

It is customary to separate as much as possible of the 
water of condensation before admitting steam to the 
cylinder. The separators used are built in many different 
shapes and types, but practically all depend upon two 
principles. These are: 

(1) Water is much more dense than steam, and if a 
stream of a mixture of water and steam be made to travel 
in a curve, the water will therefore collect at the outside 
of the curve, and 

(2) Water brought into violent contact with metallic 
surfaces " wets " them and has a tendency to adhere thereto. 

In steam separators the stream of mixture is therefore 
made to change its direction of flow suddenly and to impinge 
upon baffles in such a way that the greater part of the 
liquid is caught and drained off. 

One form of separator is shown in Fig. 231. The mixture 
impinges on sieves in the first part of its passage through 
the separator, part of the water passing through the open- 
ings and draining to the reservoir at the bottom of the 
device. Ridges and troughs catch all water separated and 
guide it to drains leading to the reservoir, so that' no water 
which is once deposited is again picked up by steam. 

Another form of separator is illustrated in Fig. 232. 
The steam impinges upon the inverted V-shaped casting 
and water caught on the projecting ridges drains toward 
the sides and then downward into the receiver, while the 
steam passes on as shown. 

162. Steam Traps. In the separators just described, 
there is a constant accumulation of water which must be 
periodically drained off if the entire device is not to fill 
up and become inoperative. Similarly there is a constant 
accumulation of liquid in steam jackets, in receivers of 
multi-expansion engines and in low points in steam lines. 



BOILER-FEED PUMPS AND OTHER AUXILIARIES 389 

To drain all such accumulations by hand from time to 
time, as necessary, would be both time consuming and 
uncertain, and devices known as traps have therefore been 
developed for doing this automatically. Traps are arranged 
to collect condensation until they fill to a predetermined 
point. When this occurs they automatically discharge in 
such a way as to prevent the escape of steam and are then 
ready to receive liquid again. 

Traps may be arranged to receive condensation from 
high or low-pressure mains or from spaces in which a vacuum 





Fig. 232.— A Steam Separator.) 



is maintained, and to discharge it to a hot well or similar 
receiver or even into the boiler itself. 

The principles upon which traps work are very nu- 
merous and there are many different designs. The more 
common either make use of the weight of the accumulating 
water to cause the trap to discharge, or they make use of 
floats resting on the condensate and opening the discharge 
valve when a certain height is reached, or they depend upon 
expansion and contraction of certain parts when exposed 
to steam of high temperature and cooler condensate respect- 
ively. 



390 STEAM POWER 

163. Steam Piping. There is a great deal of piping of 
various kinds in all steam plants and the financial success 
or failure of a plant often depends upon this apparently 
insignificant item. It is beyond the limits of a book of this 
scope to consider the many different forms of piping and the 
many different ways in which apparatus may be connected. 
This is a study in itself and one of great importance. 

It should be noted, however, that all of the following 
points must be kept in view when designing and installing 
piping and that that installation which most nearly meets 
all these requirements may be regarded as the best. 

(1) The various lines should conduct the materials flow- 
ing through them with the minimum loss of pressure and 
with the minimum loss (or gain) of heat. 

(2) The pipe lines should be so constructed as to make 
failure of a dangerous sort, from expansion and contraction, 
water hammer and such, most unlikely if not impossible. 

(3) All connections should be so made that the careless 
manipulation of valves cannot cause an accident. 

(4) The number of flange and screw connections and the 
number of valves and fittings should be reduced to the 
minimum, as they are often sources of weakness and are 
always costly. 

(5) The entire layout should be so arranged that inter- 
ruption of service because of pipe, or valve, failure is (as 
nearly as possible) impossible. 

(6) The cost of the system should be as small as it can 
be made, consistent with the other requirements. 

It is almost unnecessary to say that all of these desirable 
ends are never attained in any plant. A compromise must 
always be made in order to bring the cost within reasonable 
limits, but most of the recent installations show a tendency 
toward better design in this part of the plant and a con- 
sideration of reliability and safety far in excess of what was 
formerly customary. 



TABLES 



392 



STEAM POWER 



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SATURATED STEAM TABLE 



393 



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SATURATED STEAM TABLE 



395 



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os os o Q oi 

tH rH OS CO CO 
lO lO TjH tH rfrl 

CD CO CO CO CO 



COOiOHN 
O 00 lOCOO 
"* CO CO CO CO 

CO CO CO CO CO 



«5 N CM H O 
COCO^f CM O 
CM CM CM CM CM 

CO CO CO CO CO 



OOhpjiO 
CO CO t* CM O 



CO CO CO cO CO 



t^ OS CM CO O 
GO CO 'O CO CM 
OOOOO 
CO O co co CO 



CDOOOSrHCO TticOt^OOOs OcMCOrt^iQ CO l> 00 OS OS OhcMM^ 



OOOOO 
00 00 00 00 00 



rH r-l CO -* CO OSCMiOOSCO OO CM I> CO 00 



-^ o CO COO 
oo t^ 10 t« co' 

00 GO GO 00 00 



t^-* rH OSCO 



COrHOSt^iO OONOOON 
T^r^COCOCO CO CO CO CM CM 
0000000000 0000000000 



tO "# CM rH O 

CM CM CM CM rH 

00 00 00 00 00 



HOONCO 

rH rH OOO 

oo oo oo oo oo 



I> OS CM tO OS CO 00 CM l^ CM OOt^OI^CO ON^CMOS NiOmcMO 



rH OS 00 CO •* 
CM rH rH rH rH 

Os Os Os Cs Os 



mhooon 

rHrHrHOO 

OS OS OS OS OS 



tO T« CO rH O 

ooooo 

OS OS OS Os OS 



OS t^ CO to CO 
Os OS Os OS Os 
00 00 00 00 00 



CMrH OOSOO 

os o OS 00 CO 
oo oo co oooo 



CO "^ »0 CO CO to to CO CM O 00 to CM OS to rH t>- CO 00 CO 



CM to t^. OS CM 

to to to to CO 
CM CM CM CM CM 



^ CO 00 O CM 
co CO co t- W 
CM CM CM CM CM 



"tf CO 00 O CM 
t^ t^ t^ oo oo 

CM CM CM CM CM 



CO to t^ OOO 
00 00 00 00 OS 
CM CM CM CM CM 



CM CO to CO 00 
OS OS OS Os OS 
CM CM CM CM CM 



WON^O 


CO CM CO CO CO 


■* os ■>* co co 


00 CM COO ^ 


O0CM cooco 


Tt< lOkOcOt^- 
t^ t^ t- t^ S 


t~- CO 00 OS OS 
l> t^ t^ 1- l^ 


O O rH rH CM 

oc oooo oo oo 


CM COCOr^ tH 

go oo co co oo 


rti tO to CO CO 

oo co oo oooo 


to OS CM to t^ 


OsOOOOs 


00 t> to CO o 


00 "tf rH t^CO 


CO'* OS ^ 00 



CO »0 00 O CM "*I>OSrH(M 
CO 00 00 OS Os OS OS Os O O 
CM CM Ol CM O) CM CM CM CO CO 



-* CO 000 CM 

O O O rH rH 

CO CO CO CO CO 



C0lOl>00O rHCO^COt^ 

rH rH rH rH CM CM CM CM CM CM 

COCOCOCOCO COCOCOCOCO 



(N "<CH to 000 
iO if) IO to to 



CM <tf tO 000 

co to to to t- 



<N r* CO 000 

00 00 CO CO OS 



N^COOOO 

OS OS CD OS O 



COCOCOCOCO COCOCOCOCO COCOCOCOCO CO CO CO CO tO COCOCOCOCO 



t^OSrHCOlO t^OSrHCOtO 

COCO'*^'* «tf r* lO to »0> 



I> OS rH CO IO 

tO tOCO COO 



t^ OS rH CO IO 

cocoi>t>.t>. 



t>- OS rH CO IO 

t» t> oooo 00 



396 



STEAM POWER 





Pres. 

Abs. 

Lba. 
Sq.in. 


a 


iooiooio 

OHHNN 


oiootoo 
eoeo^Tttio 


IOOIOOIO 

io «o«o t- c- 


™>3 


* — s 


ONOCDM 
CO Ttn 00 (M CO 
(NOOONtO 


<M h ©<N <M 

IQMHHH 


<M CO tOl> O 

©00I> coco 


"* "tf< CO CO CO 


CO CO CO CO CO 


<N<M(N<N<N 


PL. 
O 
« 

w 


o 

a 

03 

> 


•e- 
< 


i-tOOO"tf O 

©oco toco 

1-H HO05 00 
HHHOO 


©OA to (MO 

OTjHt^ r-H lO 

oo rococo to 
ooooo 


© rH COtH 00 

OOCOlxN CO 
t* t* COCO <M 

©oooo 






"2 
'3 


-e- 
< 


© ^ !>• © © 

oocob- r-t to 
i-- oooo© © 


ooio<Mi>cq 

OOOHH 
T* LO »0 LO lO 


to 00 ©© © 
l>-0 CO CO © 
rH<N<N<N<N 

iO to to to to 


ooooo 


ooooo 


ooooo 


"3 

o 


g 
<1 


©<NJ1>C0© 

oo^ ot^co 

© © © 00 CO 
IQ1QIQ1Q1Q 


r^l>r^©<M 

ON^ i-i © 

oo t^ r— t^ co 

iO to to to to 


TfH © lOON 
CO CO rH © CO 

CO O CO to to 
to to to to to 






«5 

s 

p 

-< 

a 
« 

w 
n 
H 

w 

« 
I 

>< 

w 

w 


h-1 




* 


cO00©©i-i 


CO^COi>00 


©O(M<MC0 


00 00 00 00 GO 


<N(NI<M<M<N 

00 00 00 00 00 


<M CO CO CO CO 

oooooooooo 


a a 




COb- ©<M CO 


0»OON^ 


t-i 00 CO tO^H 


0005 0505 

oooor^i^t^ 


ONiONO 

© oo oo oo oo 


GO lOCO rH © 
t^ t^i>t^ CO 


13 

o 
H 




(NiOOOlMN 


CO ©CO rH^ 


OOOOOt^t^ 


OOOOt^t^t^ 

00 00 00 00 00 


<N ©)>tOCO 
t^. CO CO CO CO 

oooooooooo 


t-i 00CO^(M 
CO to to to to 

oooo oooooo 


(4 

03 




o to oco to 


coi>cotch ca 


©COfNlXM 


<N tO©<M to 
CO CO CO CO CO 


OOH^NO 
HMNNC0 
CO CO CO CO CO 


<M tOOOOCO 
COCOCOt^ t^ 
CO CO CO CO CO 


"3 
o 


.• — s 


(NOOOcOCO 


O cOiM 00 "f 


O too^JH © 


t>- 00 OOOO 
00 O0 00 00 © 


© © © © © 


-*H rt< io to to 

© © © © © 


Ph' . 


•v» 


TWOOrHCO^ 


"* CO i-i GO to 


O COO to 00 


T-H -* 00 t-H TJH 
CO CO CO i* Tt< 
COCO COCO CO 


t^-OCO to 00 
•«* to to to to 
CO CO CO CO CO 


i-H CO CO 00 O 
CO CO CO CO t^ 
co co co co co 


« 
p 

ID 

w 
« 




?^ 


IOOIOOIO 

OHHNN 


o to o to o 
CO CO ^< ^ to 


tO O to O iO 
to CO CO c- 1* 




* Gauge 
Pres- 
sure. 
Lbs. 


a 


co co coco co 


CO CO CO CO CO 


CO CO CO CO CO 


* 


©to©tOO 

©Q500H 


too to o to 

H(N<NMW 


o>oo too 

"^JH tH to to CO 



SATURATED STEAM TABLE 



397 



OiOOlOO ooooooo 

00 00 OS Ci O USOOOOOO 

»H iH iH rH C* NW^OOON 

iH <NCO 



CO00CO COO O ^h 

CO CO O rt< O tOtONCO to 

to "*i tp co cm qo to i-H o ■«* cm o 



CM <M (M CM CM T-t 



oooo 



tO-^rt< COO 

<M i— i i-H O O 
OOOOO 



8 



_ to 00 CM 

CO CM CO <M cOtO 

O O 00 00 CO t^i 



oooooo 



OOCDrtH ON 
<N tOOO'-t CO 
CO CO CO -* ■* 
tO to to tO to 



CO 00 
NNHO0 

CO 00 CO rjn CO, 
tO to COCO 1> 



OOOOO OOOOO 



CO O 00 CO CO 

lO tO Tf Tf T^ 

to to to to to 



COO 

t- CM O O _ 

<M i-H 00 b- O __ _ . 

tO tO rP ■* CO *"* ^ 



tJ< "* tO «D1> CO to 



COCO CO CO CO 

oo oooo oo oo 



rt< rJH to rH 00 O O 

oo oo oo oo i> co 



Th Tt< T^ T^ to O 00 



MOMH© 
CO CO CO CO to 



(M CO i-i 00 N O O 
rt* N O N 00 M 
NNNOiOi* 



OO 00OOCM COCOO 



o 00 CO to CO 

tO ^ TtH T^ TJH 

00 oooooooo 



CO h CO CM to O O 

nhoocooo 

00 OOt^NCO^ 



COOThir^O <Mb- 



tOOOOCN"* 
rfi -* to to to 
CO CO CO CO CO 



to (M CM 00 to - - . 
COCOrHr^ to 



CO CO!>N 00 
OOOOO 



H^MOO, 



H^OOOO i-i to 00 CO 



C0lOi> O rH 

NNNNCO 

CO CO CO CO CO 



HNrfMOO 
O i-H Tt< CO rJH t^ i 
TJH TjH Tfl T^ tO CO I 



OiOOOO 
00 00 OS O o 



ooooooo 

lOOOOOOO 



CO CO CO CO CO 



to o to o to 

CONNCO00 



CO 

CO CO CO CO • 

• ■ • • to 

to to tO to 00 

co oo oo oo o 
CMCMCO«tf * • 



o 



398 



STEAM POWER 



PROPERTIES OF ONE POUND OF SUPERHEATED STEAM 

[Condensed from Marks and Davis's Steam Tables and Diagrams, 1909, by 
permission of the publishers, Longmans, Green & Co.] 

Sp.V. = specific volume in cu.ft.; AQ = B.t.u. total heat above 
32° F.; A0 = total entropy above 32° F. 



Absolute 
Pressure. 
Lbs. Sq.in. 



Sat. Temp, 
o F> 



15 I 

213) J 



60 

(281) 



100 

(327.8) 



110 

(334.8) 



120 

(341.3) 



130 

(347.4) 



140 

(353.1) 



150 

(358.5) 



Sp. V. 
AQ 
Acj> 

Sp. V. 
AQ 

A0 

Sp. V. 
AQ 

A0 

Sp. V. 
AQ 

A0 

Sp.V. 
AQ 
Acf> 

Sp. V. 
AQ 
A<f> 

Sp. V. 
AQ 
A0 

Sp. V. 
AQ 
A(f> 



Degrees of Superheat. 



26.27 
1150.7 
1 . 7549 

8.51 
1173.6 
1.6581 

4.43 
1186.3 
1 . 6020 

4.05 
1188.0 
1 . 5942 

3.73 
1189.6 
1.5873 

3.45 
1191.0 
1 . 5807 

3.22 
1192.2 
1.5747 

3.01 
1193.4 
1.5692 



w 



28.40 
1174.2 
1.7886 

9.19 
1198.8 
1 . 6909 

4.79 
1213.8 
1.6358 

4.38 
1215.9 
1 . 6282 

4.04 
1217.9 
1.6216 

3.74 
1219.7 
1.6153 

3.49 
1221.4 
1 . 6096 

3.27 
1223.0 
1.6043 



100 



30.46 
1197.6 
1.8199 

9.84 
1223.4 
1.7211 

5.14 
1239.7 
1.6658 

4.70 
1242.0 
1 . 6583 

4.33 
1244.1 
1.6517 

4.02 
1246.1 
1.6453 

3.75 
1248.0 
1.6395 

3.51 
1249.6 
1.6343 



150 



32.50 
1221.0 

1.8492 

10.48 
1247.7 
1.7491 

5.47 
1264.7 
1 . 6933 

5.01 
1267.1 
1.6857 

4.62 
1269.3 
1.6789 

4.28 
1271.4 
1.6724 

4.00 
1273.3 
1.6666 

3.75 
1275.1 
1 . 6612 



200 



34.53 
1244.4 

1.8768 

11.11 
1271.8 
1.7755 

5.80 
1289.4 

1.7188 

5.31 
1291.9 
1.7110 

4.89 
1294.1 
1.7041 

4.54 
1296.2 
1.6976 

4.24 

1298.2 
1.6916 

3.97 
1300.0 
1.6862 



250 



36.56 
1267.7 
1.9029 

11.74 
1295.8 
1.8002 

6.12 
1313.6 
1.7428 

5.61 
1316.2 
1 . 7350 

5.17 
1318.4 
1 . 7280 

4.80 
1320.6 
1.7213 

4.48 
1322.6 
1.7152 

4.19 
1324.5 
1.7097 



300 



38.58 
1291.1 
1.9276 

12.36 
1319.7 
1.8237 

6.44 
1337.8 
1.7656 

5.90 
1340.4 
1.7576 

5.44 

1342.7 
1.7505 

5.05 
1344.9 
1.7437 

4.71 
1346.9 
1 . 7376 

4.41 

1348.8 
1 . 7320 



SUPERHEATED STEAM TABLE 



399 



PROPERTIES OF ONE POUND OF SUPERHEATED STEAM 

(Continued) 



Absolute 
Pressure. 
Lbs. Sq.in. 



Sat. Temp. 

o F _ 



160 

(363.6) 



170 

(368.5) 



180 

(373.1) 



190 

(377.6) 



200 

(381.9) 



300 

(417.5) 



500 

(467 . 3) 



Degrees of Superheat. 



Sp. V. 
AQ 

A<t> 

Sp. V. 
AQ 

Acf> 

Sp. V. 
AQ 
A<£ 

Sp. V. 
AQ 
A0 

Sp.V. 
AQ 
A</> 

Sp.V. 
AQ 
A0 

Sp. V. 
AQ 
A</> 



2.83 
1194.5 
1.5693 

2.68 
1195.4 
1.5590 



50 



3.07 
1224.5 
1 . 5993 



2.91 
1225.9 
1.5947 



100 



2.53 2.75 
1196.4 1227.2 
1.5543 1.5904 



2.41 
1197.3 
1.5498 

2.29 
1198.1 
1.5456 

1.55 
1204.1 
1.5129 

0.93 

1210.0 

1.470 



3.30 
1251.3 
1 . 6292 



3.12 
1252.8 



150 



3.53 
1276.8 
1.6561 

3.34 

1278.4 



200 



1.6246 1.6513 



2.96 
1254.3 
1.6201 



2.62 
1228.6 
1.5862 

2.49 
1229.8 
1 . 5823 

1.69 
1240.3 
1.5530 

1.03 

1256 

1.519 



2.81 
1255.7 
1.6159 



2.68 
1257 
1.6120 



1.83 
1268 



3.16 

1279.9 
1.6468 



3.00 
1281.3 
1.6425 



3.74 
1301.7 
1.6810 

3.54 
1303.3 
1 . 6762 

3.35 
1304.8 
1.6716 

3.19 
1306.3 
1.6627 



250 



300 



2.86 3.04 
1282.6 1307.7 
1 . 6385 1 . 663? 



1.96 
1294.0 



1.5824 1.6082 



1.11 

1285 
1.548 



1.22 

1311 

1.573 



2.09 
1319.3 
1.6323 

1.31 

1337 

1.597 



3.95 
1326.2 
1.7043 

3.73 
1327.9 
1.6994 

3.54 
1329.5 
1.6948 

3.37 
1330.9 
1 . 6904 

3.21 
1332.4 
1.6862 

2.21 

1344.3 
1 . 6550 



4.15 
1350.6 
1 . 7266 

3.92 
1352.3 
1.7217 

3.72 
1353.9 
1.7169 

3.55 
1355.5 
1.7124 

3.38 
1357.0 

1.7082 

2.33 
1369.2 
1.6765 



1.39 

1362 

1.619 



1.47 

1388 

1.640 



INDEX 



PAGE 

Absolute pressures 41 

Absolute temperature scale 12 

Action of steam, in cylinder 24 

on impulse blades of steam turbine 234-236 

Adiabatic expansion 58 

Advance angle 167 

Advantages of condensing 251, 252 

Advantages, relative, of contact and non-contact condensers. . 274, 275 

Air, excess, combustion 286 

Advantages and disadvantages of 311 

Analogy, hydraulic 26 

Analyses of coal (see fuels) 299-301 

Purchase of coal on analysis 302 

Angle of advance 167 

Ash in coal 300 

Atmospheric line on indicator diagram 119 

Atoms 278 

Avogadro's Law 281 

Babcock & Wilcox superheater 367 

water-tube boiler 352-355 

Balanced slide valves 184 

Barometer, conversion of readings from inches mercury to pounds 

per square inch 255, 256 

Barometric Condenser 261-266 

Baume scale to express gravity 303 

Bearings Ill, 112 

Bilgram diagram 168-182 

Angularity of connecting rod 179 

Diagram for both cylinder ends 177 

Exhaust and compression 175-177 

Indicator diagram from 180-183 

Piston positions 177-182 

Blades, impulse, action of steam on, in impulse turbine 234-236 

401 



402 INDEX 



PAGE 



Boiler-feed pumps and other auxiliaries 382-390 

Boiler, generation of steam in 38, 39 

Boilers, steam 305-374 

Circulation in 341, 342 

Classification according to — 

(1) form; (2) location of furnace; (3) use; (4) direction 
of principal axis ; (5) relative positions of water and hot 

gases 305, 306 

Draft apparatus 370-373 

Chimneys or stacks 370-372 

Mechanical draft 372, 373 

Effects of soot and scale 364, 365 

Efficiencies 363, 364 

Functions of parts 306-308 

Furnaces and combustion 308-311 

Hand firing 311-315 

Mechanical grates 315, 316 

Mechanical stokers 317-335 

Rate of combustion 335-337 

Rating 358-362 

Boiler horse-power 359 

Equivalent evaporation 361 

Scale 365, 366 

Prevention of 366 

Smoke and its prevention 316, 317 

Strength and safety 337-341 

Superheaters — 

Built in 366, 367 

Separately fired 366, 367 

Babcock & Wilcox 367 

Foster 369, 370 

Heine 367, 368 

Types of boilers 342-359 

Babcock & Wilcox, water tube 352-355 

Continental 346-348 

Externally fired, return tubular 349-352 

Heine water-tube 355, 356 

Internally fired, tubular 342-345 

Locomotive 346 

Scotch marine 348, 349 

Sterling water-tube 356-358 

Vertical fire tube 343-345 

Wickes vertical water-tube 358, 359 



INDEX 403 



TAGE 



British Thermal unit 3, 13 

Buildings, heating of, by exhaust steam 376, 377 

Built-in superheaters 366, 367 

Calorific value of coals — 

Dulong's formula 301, 302 

Fuel Calorimeter 302 

Calorific value of petroleum oils 303, 304 

Calorimeter, fuel 302 

Carbon, combustion of 279 

CO, combustion to 279-282 

C0 2 combustion to 282, 283 

CO and C0 2 , conditions determining formation of. . . 284-286 

CO to C0 2 , combustion of 283, 284 

complete combustion of 308-310 

flue gases from combustion of 286, 287 

Card factors and conventional diagram 125-128 

Centigrade scale 10> H> * 2 

Chain grate stokers 318-322 

Chart— 

Mollier, for steam 230 

temperature-entropy, for steam 62-65 

Chimneys or stacks 370-372 

Circulation in boilers 341-342 

Classification of boilers, according to — 

(1) form; (2) location of furnace; (3) use; (4) direction of prin- 
cipal axis; (5) relative postion of water and hot gases 305-374 

Classification of steam engines 92, 93 

Clearance — steam engine — 

mechanical and volumetric 84, 85 

Clearance volume determined from diagram 131, 132 

Closed and open feed-water heaters 377-380 

Coal-fuels , 297-299 

Analyses of — proximate and ultimate 299-301 

Purchase of, on analysis 302 

Coefficient, excess in combustion 286 

Combined indicator diagrams 155-158 

Combined type turbine 247 

Combustion and furnaces; steam boilers 308-311 

Combustion 277-295 

Definitions — Compounds, elements, heat or calorific value, 
atoms, molecules, etc 277-279 



404 INDEX 

PAGE 

Combustion — 
Combustion of — 

Carbon 279 

Hydrocarbons 289, 290 

Hydrogen 287-289 

Calorific value of 290 

Mixtures 290, 291 

Sulphur 290 

Combustion to — 

CO 279-282 

C0 2 282, 283 

CO to C0 2 283, 284 

Conditions determining formation of CO and C0 2 284-286 

Excess Air and excess coefficient 286 

Flue gases from combustion of carbon 286, 287 

Rate of, in boiler furnaces 335-337 

Temperature of combustion 291-294 

Theoretical temperature 292 

Commercial fuels — solid, liquid and gaseous 296, 297 

Complete expansion cycle 55-58, 72 

Complete TV-chart for steam 68-70 

Compound engine 149-151 

Compounding 141-158 

Combined indicator diagram 155-158 

Compounding 144-149 

Cylinder ratios 151-153 

Gain by expansion 141-144 

Indicator diagrams and mean pressures 153-155 

The compound engine 149-151 

Compounds — combustion 277-279 

Compression and exhaust — Bilgram diagram 175-177 

Condensation, cylinder, methods of decreasing 89-92 

Condensation, initial , 81, 82 

determination of 86-89 

Condensers and related apparatus 251-276 

Advantages of condensing 251, 252 

Conversion of readings from inches of mercury to 

lbs. per square inch 255, 256 

Cooling towers 275, 276 

Measurement of vacuum 252-255 

Principle of 256-258 

Types of — 

Contact 258-268 



INDEX 405 

PAGE 

Condensers (continued)— 
Contact — 

Barometric 261-266 

Jet, Parallel flow 259, 260 

Siphon 266 

Westinghouse— Leblanc 267, 268 

Non-contact 268-271 

Surface 268-270 

Two-pass or double flow 270-271 

Relative advantages 274, 275 

Water required by contact condensers 271-273 

Water required by non-contact condensers "273, 274 

Condensing, advantages of 251, 252 

Condensing plants 23 

Conditions determining formation of CO and C0 2 284-286 

Connecting rod 109, 110 

Angularity of 179 

Conservation of Energy, law of 2 

Conservation of Matter, law of 1 

Constant-quality lines on T^-chart 66, 67 

Constant volume lines, on T$-chart 68 

Constant speed governing 215, 216 

Contact condensers 258-268 

Continental type boiler .' 346-348 

Conventional diagram and card factors 125-128 

Conversion of barometric readings, from inches mercury to pounds 

per square inch 255, 256 

Cooling towers 275, 276 

Corliss and other high-efficiency engines 196-212 

Locomobile type 210-212 

Non-detaching Corliss gears 201-205 

Poppet valves 205-208 

Trip-cut-off Corliss 196-201 

Unaflow engine 208-210 

Corliss engine, trip-cut-off 196-201 

Corliss gears, non-detaching 201-205 

Crank end of engines 98 

Cross-head and guides 107, 108 

Cushion steam and cylinder feed 85, 86 

Cut-off governing 215 

Cut-off ratio 128 

Cycle, area on 7'</>-chart representative of work 73 

Complete expansion 55-58, 72 



406 INDEX 

PAGE 

Cycle, incomplete expansion 58-60, 74, 75 

Modifications for wet and superheated steam 73, 74 

Of events in simple steam power plant 22 

Theoretical, of steam turbine 225-228 

Cycles, desirability of various, in engines 55 

Cylinder, action of steam in 24 

Condensation, methods of decreasing 89-92 

Efficiency 139 

Feed and cushion steam 85, 86 

Ratios 151-153 

Cylinder and steam chest 101, 102 

Decreasing cylinder condensation 89-92 

De Laval impulse turbine 236-238 

Density, specific, of dry saturated steam 38 

Description and method of operation of D-slide valve 159-165 

Design of nozzle, steam turbine 228-234 

Determination of clearance volume from diagram 131, 132 

Determination of I.h.p 120-124 

Developed horse-power 137 

Developed thermal efficiency 138 

D-slide valve 159-195 

Angle of advance 167 

Angularity of connecting rod ' 179 

Bilgram diagram 169 

Description and method of operation 159-165 

Diagram for both cylinder ends 177 

Exhaust and compression 175 

Exhaust lap 168 

Indicator diagram from Bilgram diagram 180 

Lead 166, 167 

Limitations of D-slide valve 183-185 

Piston positions 177 

Reversing engines , 185-187 

Steam lap — outside lap 165, 166 

Valve setting 187-195 

D-slide valve engine, simple 96-98 

Diagram, Bilgram, for both cylinder ends 177 

Bilgram, indicator diagram from 180-183 

Indicator 24 

Indicator and mean pressures for compound engines . 153-155 

combined 155-158 

Indicator, conventional and card factors 125-128 



INDEX 407 

PAGE 

Diagram, water rate 86, 132-136 

Diagrams from real engine 192, 193 

Double acting engines 55" 

Double-flow condenser 2 70 

Downdraft furnace 315 

Draft apparatus 370-373 

Chimneys or stacks 370-372 

Mechanical draft 372, 373 

Dry-air pump 264 

Dry-saturated steam, total heat of 33 

Specific density of 38 

Specific volume of 36-38 

Dry-vacuum pump 264 

Dulong's formula — combustion 301, 302 

Duplex steam pump 383, 384 

Eccentric 160-165 

Economy of turbines 247-249 

Effective pressure, mean, methods of varying 215 

Efficiencv 52, 53 

Cylinder 139 

Developed thermal 138 

Effect of temperature range on 75 

Indicated thermal 138 

Mechanical and thermal 137-140 

Of boilers 363, 364 

Relative 139 

Elements — combustion 277 

Energy — 

Conservation of energy, law of 2 

Heat 2 

Mechanical 2 

Units of 3 

Engine — 

Application of theory for an ideal to a real 54, 55 

Compound, triple, quadruple, quintuple 148 

Receiver type 149 

Tandem and cross-compound 151 

Woolf type 149 

Desirability of various cycles 55 

Double acting 55 

Efficiency 52, 53 

Heat quantities involved 50-52 



408 INDEX 



PAGE 



Engine (continued) — 

Ideal steam 43-60 

Operation of 45, 42 

Operation of the real steam 77-80 

Reversing 185-187 

Steam — 

Classification — 

1) On basis of rotative speed; (2) Ratio of 
stroke to diameter; (3) Valve gear; (4) 
Position of longitudinal axis; (5) Num- 
ber of cylinders; (6) Cylinder arrange- 
ment; (7) Use 92, 93 

Clearance, volumetric and mechanical 84, 85 

Crosshead and guides 107, 108 

Cushion steam and cylinder feed 85, 86 

Diagram water rate 86 

Cylinder and steam chest 101, 102 

Determination of initial condensation 86-89 

Initial condensation 81 

Losses, in real installations 80-84 

Methods of decreasing cylinder condensation. .. 89-92 

Nomenclature 98 

Principal parts 98-114 

Bearings Ill, 112 

Connecting rod 109, 110 

Crosshead and guides 107, 108 

Cylinder and steam chest 101, 102 

Flywheels 112, 113 

Frame 99, 100 

Piston. 102-106 

Piston rod and tail rod 106, 107 

Shaft 110, 111 

Re-evaporation in 83, 84 

Rotative and piston speed 93-96 

Simple. D-slide valve 96-98 

Throttling or wire-drawing 82, 83 

Work done by 46-60 

Engines, Corliss and other high efficiency 185-187 

Locomobile type 210-212 

Non-detaching Corliss gears 201-205 

Poppet valves 205-208 

Trip-cut-off Corliss 196-201 

Unaflow 208-210 



INDEX 409 



PACE 



Entropy diagram 61-71 

of liquid, vaporization, and dry saturated steam 01-63 

T0-chart for steam 62-65 

Complete 68-70 

Constant quality lines 66, 67 

Diagram for a real engine 136 

Heat from 68 

Quality from 65-68 

Saturation curve 63 

Superheating lines 63, 64 

Volume from 68 

Water line 63 

Entropy, diagrams of steam cycles 72-76 

Equivalent evaporation, boilers 361 

Excess air — combustion 286 

Advantages and disadvantages 311 

Excess coefficient 286 

Exhaust and compression — Bilgram diagram 175-177 

Exhaust lap 166, 168 

Exhaust steam, utilization of, for heating buildings 376, 377 

Expansion, adiabatic 58 

Cycle, the complete 55-58, 72 

the incomplete 58-60 

Gain by, in compounding 151-144 

Ratio of, apparent and real 128-130 

External latent heat of vaporization 31 

Externally fired, return tubular boiler 349-352 

Fahrenheit scale 11, 12 

Feed-water heating 377 

Open and closed heaters 377-380 

Firing boilers by hand 311-315 

Fixed carbon in coal 300 

Flywheel 97, 98, 112, 113 

Regulation 213, 214 

Flue gases from combustion of carbon 286, 287 

Foot-pound, definition 3 

Forward stroke of engines 98 

Foster superheater 369, 370 

Frames of engines 99, 100 

Front end of engines 98 

Fuel calorimeter 302 

Fuels 296-304 



410 INDEX 



PAGE 



Fuels (continued) — 
Commercial — 

Solid, liquid, gaseous 296, 297 

Coal - • • 297-299 

Analyses 299-301 

Calorific value of — 

Dulong's formula 301, 302 

Fuel Calorimeter 302 

Petroleum 302 

Baume scale to express gravity 303 

Calorific values 303, 304 

Purchase of coal on analysis 302 

Functions of boiler parts 306-308 

Furnaces — and combustion 308-311 

— Updraft and downdraft -. . 315 

Gases, and vapors, steam 27 

— Flue, from combustion of carbon 286, 287 

Gaseous fuels , 296, 297 

Gauge pressure 39-41 

Gearing and staging — turbines 238-243 

Gears, Corliss, non-detaching 201-205 

Generation of steam in real steam boiler 38, 39 

Generation of steam or water vapor 28 

Governing — throttle and cut-off 215 

Coefficient of regulation 216 

Constant speed 215, 216 

Governor 97, 98 

Regulation 214, 215 

Governors — 

Pendulum 217 

Rites inertia 218-220 

Shaft 217, 218 

Grates, mechanical 315, 316 

Gridiron valve 185 

Guides and crosshead 107, 108 

Hand firing — steam boilers 311-315 

Heat 9 

Absorption, reversal of process 38 

Energy 2 

Unit of 13 

From 7>-chart 68 



INDEX 411 

PAGE 

Heat (continued) — 

Latent, of vaporization 30, 32 

Internal and external 30, 31 

Of liquid, q or h 31, 32 

Of superheat 34, 35 

Quantities in rectangular cycle 50-52 

Quantity of 16 

Specific 14 

Total, of dry saturated steam 33 

Of superheated steam 36 

Of wet steam 33, 34 

Value of elements and compounds 277-279 

Heat, waste — in steam plant 375-381 

Feed-water heating 377 

Open and closed heaters 377-380 

Utilization of exhaust for heating buildings 376, 377 

Heaters, feed-water, open and closed 377-380 

Heine superheater 367, 368 

Heine water-tube boiler 355, 356 

Horizontal, return, tubular boiler 306-308 

Horse-power 1 ' 

Developed 137 

Hour, definition 18 

Of steam boilers 359 

Hydraulic analogy 26 

Hydrocarbons, combustion of 289, 290, 310 

Calorific value of 290 

Hydrogen, combustion of 287-289 

I.h.p. — determination of 120-124 

Impulse steam turbine 221-225 

De Laval type 236-238 

Inclined stokers 322-325 

Incomplete expansion cycle 58-60, 74, 75 

Indicated thermal efficiency 138 

Indicator H5 

Indicator diagram 24, 115-140 

Atmospheric line 119 

Conventional and card factors 125-129 

Cut-off ratio 128 

Determination of clearance volume from diagram. 131, 132 

Determination of I.h.p 120-124 

Diagram factor or card factor 126-129 



412 INDEX 



PAGE 



Indicator diagram (continued) — 

From Bilgram diagram 180-183 

Mean effective pressure 122 

Planimeter 123 

Ratio of expansion . . . 128-131 

Reducing mechanism 118 

Scale of spring 118 

Indicator diagrams and mean pressures for compound engines 153-155 

Combined 155-158 

Indicator diagrams from real engine * 192, 193 

Injector, steam e 384-386 

Inertia governor, Rites 218-220 

Initial condensation 81, 82 

Determination of 86-89 

Inside lap, negative 168 

Internal latent heat of vaporization , 30 

Internally fired, tubular boilers 342-345 

Jet condensers 259, 260 

Joule's equivalent 14 

Joule, the 3 

Kinetic mechanical energy 8 

Lap angle 166 

Lap, steam 165, 166 

Negative inside 168 

Outside and exhaust 166, 168 

Latent heat of vaporization 30, 32 

Internal and external 30, 31 

Lead 166, 167 

Leblanc — Westinghouse condenser 267, 268 

Liquid fuels 296, 297 

Liquid, heat of, q or h 31, 32 

Entropy of 61 

Limitations of D-slide valve 183-185 

Balanced slide valves 184 

Gridiron valve 185 

Piston valve 184 

Riding cut-off valves 185 

Locomobile type of high efficiency engines 210-212 

Locomotive type boiler 346 

Low-pressure or exhaust steam turbines 249 



INDEX 413 

PAGE 

Matter 1 

Law of conservation of matter 1 

Units of matter 3 

Mean effective pressure 122 

Methods of varying 215 

Mean pressures and indicator diagrams for compound engines . 153-155 

Measurement of temperature 10 

Measurement of vacuums 252-255 

Mechanical and thermal efficiencies 137-140 

Mechanical clearance, steam engine 84, 85 

Mechanical draft. 372, 373 

Mechanical energy 2, 3, 7 

Potential and kinetic 7, 8 

Mechanical grates 315, 316 

Mechanical stokers 317-335 

Mercury readings, conversion to pounds per square inch 255, 256 

Mercury thermometers 10-12 

Method of operation and description of D-slide valve 159-165 

Mixtures, combustion of 290, 291 

Moisture in coal 299 

Molecular activity 9 

Molecules 278 

Natural draft, chimneys 370 

Negative inside lap 168 

Non-condensing plants 23 

Non-contact condensers 2t3-271 

Surface (Wheeler) 268-270 

Non-detaching Corliss gears 201-205 

Nozzle design, steam turbine 228-234 

Oil firing 333-335 

Open and closed feed-water heaters 377-380 

Operation of simplified steam engine 45, 46 

Operation of real steam engine 77-80 

Outside steam lap 166 

Outstroke of engine 98 

Parallel-flow condenser 259-261 

Parson's type turbine 246 

Pendulum governors 217 

Petroleum 302 

Baume scale to express gravity of 303 

Calorific values 303, 304 



414 INDEX 

PAGE 

Piping, steam 390 

Piston, engine 102-106 

Piston positions for Bilgram diagram 177-182 

Piston rod and tail rod 106, 107 

Piston speeds of steam engines 93-96 

Piston valve 184 

Planimeter 123 

Plant, steam power 20 

Plants, condensing, non-condensing 23 

Poppet valves 205-208 

Positions of piston for Bilgram diagram 177-182 

Potential mechanical energy 7 

Powdered coal stokers 333 

Power and work 17 

Power, unit of, horse power 17 

Pressure, absolute 41 

Gauge 39-41 

Mean effective 122 

Methods of varying 215 

Pressures, mean, and indicator diagrams for compound engines. 153-155 

Prevention of smoke 316, 317 

Prime-mover 20 

Principal parts of engines 98-114 

Principle of condenser 256-258 

Properties of steam 27 

Proximate analysis of coal 299 

Pump, dry air or dry vacuum 264 

Vacuum 259 

Pumps, boiler feed 382-384 

Purchase of coal on analysis 302 

Quality from '" ^-chart 65-68 

Constant, lines 66, 67 

Quantity of heat 16 

Rate, diagram water 86, 132-136 

Rate of combustion in boiler furnaces 335-337 

Rating of steam boiler 358-362 

Ratio, cut-off 128 

Ratio of expansion — apparent and real 128-130 

Ratios, cylinder 151-153 

Reaction type turbine 243-247 

Receiver engine 149 



INDEX 415 

p \av. 

Recovery of waste heat 375-381 

Reducing mechanism 118 

Re-evaporation 83 

Regulation 213-220 

Coefficient of governor 216 

Constant speed governing 215, 216 

Governors — 

Pendulum 217 

Rites inertia 218-220 

Shaft 217,218 

Kinds — flywheel and governor 213-215 

Methods of varying mean effective pressure — Throt- 
tling and cut-off 215 

Relative advantages of contact and non-contact condensers. . 274, 275 

Relative efficiency 139 

Return tubular boilers, horizontal 306-308, 349-352 

Reversal of process of heat absorption 38 

Reversing engines 185-187 

Riding cut-off valve 185 

Rites inertia governor 218-220 

Rotative speeds of steam engines 93-96 

Safety and strength of boilers 337-341 

Saturated steam, dry, specific volume of 36-38 

Saturated vapor 31 

Saturation curve, temperature entropy chart for steam 63 

for compound engine cards 156 

Scale 365, 366 

Prevention of 366 

Scale of spring, indicator 118 

Scotch marine type boiler 348, 349 

Separately fired superheaters 366, 367 

Separators 386-388 

Setting, valve 187-195 

Shaft governors 217, 218 

Shaft of engine 110, 111 

Simple D-slide valve engine 96-98 

Siphon condensers 266 

Slide valves 184 

Balanced 184 

Gridiron valve 185 

Piston valve 184 

Riding cut-off valve 185 



416 INDEX 

PAGE 

Smoke and its prevention 316, 317 

Solid fuels 296, 297 

Soot and scale, effects of, in boilers 364, 365 

Specific density of dry, saturated steam 38 

Specific heat 14 

Specific volume of dry saturated steam 36-38 

Speeds, rotative and piston, of steam engines 93-96 

Spring, scale of, indicator 118 

Stacks or chimneys 370-372 

Staging and gearing, steam turbines 238-243 

Steam, action in cylinder 24 

Action of, on impulse blades of turbine 234-236 

Boiler, generation of steam in 38, 39 

Cushion, and cylinder feed 85, 86 

Diagram water rate 86 

Cycles, T</>-diagrams of, 72-76 

Steam engine the ideal 43-60 

Bearings, Ill, 112 

Classification 92, 93 

Connecting rod 109, 110 

Crosshead and guides . v 107, 108 

Cylinder and steam chest 101, 102 

Determination of initial condensation 86-89 

Flywheel and governor 97, 98 

Flywheels 112, 113 

Frame 99, 100 

Losses in real installations 80-84 

The real 77-114 

Initial condensation 81 

Re-evaporation 83, 84 

Throttling 82 

Wire-drawing 82, 83 

Methods of decreasing cylinder condensation .... 89-92 

Nomenclature of 98 

Operation of, 77-80 

Piston 102-106 

Piston rod and tail rod 106, 107 

Principal parts 98-114 

Rotative and piston speeds 93-96 

Simple D-slide valve 96-98 

Steam, entropy of dry saturated 61, 62 

Generation of 28-39 

Heat of superheat 34-35 



INDEX 417 

PAG E 

Steam, lap, D-slide valve 165, 166 

Modification of T^-chart for wet and superheated 73, 74 

Properties of 27 

Specific density of dry saturated 38 

Specific volume of dry saturated 36-38 

Temperature-entropy chart for 62-71 

7>-chart complete 68-70 

Total heat of dry saturated 33 

Total heat of wet 33, 34 

Vapors and gases 27 

Wet, effect of 53 

Steam injector 384 

Steam piping 390 

Steam power plant 20-22 

Steam trap 388 

Steam turbine (see Turbine) 

Stephenson link gear , 186 

Sterling water-tube boiler 356-358 

Stokers, mechanical 317-313 

Chain grate 318 

Inclined, oVerfeed 323-327 

Powdered coal 333 

Sprinkler 318 

Underfeed 327-333 

Strength and safety of boilers 337-341 

Sulphur, combustion of 290 

Sulphur in coal 301 

Superheat, heat of 34, 35 

total heat of 36 

Superheaters — 

Built in 366, 367 

Separately fired 366, 367 

Babcock and Wilcox 367 

Foster 369, 370 

Heine 367, 368 

Superheating 31 

Lines, on temperature-entropy chart for steam . . 63, 64 
Surface condensers 268, 269 

Tail rod and piston rod of engine 106, 107 

Temperature 9 

Measurement of 10 

Pressure relations 29 



418 INDEX 

PAGE 

Temperature of combustion 291-294 

Temperature rise 293 

Theoretical 292 

Temperature-entropy chart for steam 62-71 

Complete chart 68-70 

Heat from 68 

Quality from 65-68 

Volume from 68 

T<£-diagram for a real engine 136 

T ^-diagrams of steam" cycles 72-76 

Complete expansion cycle 72 

Area of cycle representative of work 73 

Modifications for wet and superheated steam 73 

Temperature range, effect on efficiency 75 

Temperatures of vaporization 29 

Theoretical cycle of steam turbine 225-228 

Thermal and mechanical efficiency 137-140 

Developed thermal efficiency 138 

Indicated thermal efficiency 138 

Thermometers, mercury 10-12 

Throttle governing 215 

Throttling or wire-drawing 82 

Towers, cooling 275, 276 

Traps, steam 388, 389 

Trip-cut-off Corliss engine 196-201 

Triple expansion 148 

Tubular boiler, horizontal return 306-308 

Turbine, steam 221-250 

Action of steam on impulse blades 234-236 

Combined type 247 

De Laval impulse type 236-238 

Economy of 247-249 

Gearing and staging 238-243 

Impulse 221-225 

Nozzle design 228-234 

Reaction type 243-247 

Theoretical cycle 225-228 

Types of boilers 342-359 

Babcock & Wilcox, water-tube 352-355 

Continental 346-348 

Externally fired, return tubular 349-352 

Heine water-tube 355, 356 

Internally fired, tubular 342-345 



INDEX 419 

PAGE 

Types of boilers (continued) — 

Locomotive 346 

Scotch marine 348, 349 

Sterling water-tube 356-358 

Wickes vertical water-tube 358, 359 

Types of condensers — 

Contact 258-268 

Barometric 261-266 

Jet, parallel flow type 259, 260 

Siphon 266 

Westinghouse-Leblanc 267, 268 

Non-contact — 

Surface , 268-270 

Two-pass or double flow 270, 271 

Ultimate analysis of coal 299-301 

Una-flow engine 208-210 

Underfeed stokers 325-333 

Unit of heat energy 13 

Units of matter, energy and work 3 

Updraft furnace 315 

Utilization of exhaust steam for heating buildings 376, 377 

Vacuum 252-255 

Measurement of 253, 254 

Pump 259 

Valve, D-slide — (see D-slide valve) 159-195 

Setting 187-195 

Valves 159-195, 205-208 

Balanced slide 184 

Gridiron 185 

Piston 184 

Poppet 205-208 

Riding cut-off 185 

Vapor, saturated 31 

Pressure-temperature relations; saturated water vapor. ... 29 
Water or steam, generation of 28 

Vaporization, entropy of 61 

Latent heat of 30, 32 

Internal and external 30, 31 

Temperatures of 29 

Vapors and gases 27 

Volatile matter in coal 299 



420 INDEX 

PAGR 

Volume, clearance, determined from diagram 131, 132 

Constant, line 68 

From T</>-chart 68 

Volumetric clearance, steam engine 84, 85 

Waste heat, in steam plant 375, 376 

Recovery of 375-381 

Water line, temperature entropy chart for steam 63 

Water rate, diagram, steam engine 86, 132-136 

Water required by contact condensers 271-273 

Water-tube boilers 352-359 

Water vapor or steam, generation of 28 

Water vapor, saturated, pressure- temperature relations 29 

Weight of water required by non-contact condensers 273, 274 

Westinghouse-Leblanc condenser ' 267, 268 

Westinghouse-Parsons turbine 245 

Wet and superheated steam; modifications of 7'0-chart for. ... 73, 74 

Wet steam, total heat of 33, 34 

Wickes vertical water-tube boiler 358, 359 

Wire-drawing or throttling 82 

Woolf type engine 149 

Work 3, 17 

Area of cycle on 7 1 <£-diagram representative of work 73 

Done by the engine 46-50 

Unit of 3 



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